Vehicle control device, and vehicle control method

ABSTRACT

A vehicle control device includes a sensor and a controller. The sensor detects wheel speed. The controller estimates sprung mass state based on detected information in a prescribed frequency range. The controller controls a variable-damping force shock absorber to bring the estimated sprung mass state to a target sprung mass state. The controller estimates wheel rim braking/drive torque acting on a wheel. The controller determines the estimation accuracy of the sprung mass state has deteriorated when a rate of change of a stationary component extracted from components of wheel rim braking/drive torque acting on a wheel is detected to equal or exceed a prescribed value. The controller controls the variable-damping force shock absorber to a more limited extent than when the estimation accuracy has not deteriorated.

CROSS-REFERENCE TO RELATED APPLICATIONS

This application is a U.S. National stage application of InternationalApplication No. PCT/JP2013/063240, filed May 13, 2013, which claimspriority to Japanese Patent Application No. 2012-110303 filed in Japanon May 14, 2012 and Japanese Patent Application No. 2012-110304 filed inJapan on May 14, 2012.

BACKGROUND

1. Field of the Invention

The present invention relates to a control device and a control methodfor controlling the state of a vehicle.

2. Background Information

Japanese Laid-Open Patent Application No. 2009-241813 discloses atechnology for estimating stroke speed from fluctuations in wheel speedin a predetermined frequency region, and modifying the damping force ofa variable damping force shock absorber in accordance with the strokespeed, to control the sprung behavior.

SUMMARY

However, a problem encountered in the aforementioned prior art is thatwhen a disturbance of wheel speed arises within the predeterminedfrequency region, the accuracy of estimation of the stroke speed islowered. With the foregoing in view, it is an object of the presentinvention to provide a control device and a method for controlling avehicle, whereby it is possible to control the vehicle body orientation,even when the accuracy of estimation of the sprung mass state islowered.

To achieve the aforementioned object, the vehicle control device of thepresent invention estimates a sprung mass state based on wheel speedinformation in a predetermined frequency region, and controls a variabledamping force shock absorber so as to bring the estimated sprung massstate to a target sprung mass state. At this time, when low accuracy ofestimation in the sprung mass state has been detected, the damping forceof the variable damping force shock absorber is transitioned to a fixeddamping force corresponding to a vehicle state quantity existing priorto lowering of the accuracy of estimation.

Therefore, it can be detected when low accuracy of estimation in thesprung mass state has arisen, and situations in which control continueswhile the accuracy of estimation remains lower can be avoided. Moreover,situations in which control is performed incorrectly due to the limiteddamping force control performed when the accuracy of estimation is lowcan be minimized, and a stable vehicle body orientation can be achieved.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a schematic system diagram showing a vehicle control deviceaccording to a first embodiment.

FIG. 2 is a control block diagram showing a configuration of controlperformed by the vehicle control device according to the firstembodiment.

FIG. 3 is a conceptual diagram showing a configuration of a wheel speedfeedback control system according to the first embodiment.

FIG. 4 is a control block diagram showing the configuration of a drivingstate estimating unit of the first embodiment.

FIG. 5 is a control block diagram showing the specifics of control in astroke speed calculator unit of the first embodiment.

FIG. 6 is a block diagram showing the configuration of a reference wheelspeed calculator unit of the first embodiment.

FIG. 7 is a schematic diagram showing a vehicle body vibration model.

FIG. 8 is a control block diagram showing brake pitch control in thefirst embodiment.

FIG. 9 is a graph simultaneously showing a wheel speed frequency profiledetected by a wheel speed sensor, and a stroke frequency profile from astroke sensor not installed in the present embodiment.

FIG. 10 is a control block diagram showing frequency-sensitive controlin sprung mass vibration damping control in the first embodiment.

FIG. 11 is a correlation graph showing human sensation profiles indifferent frequency regions.

FIG. 12 is a plot showing the relationship between the proportion ofvibration contamination and damping force in a float region in thefrequency-sensitive control of the first embodiment.

FIG. 13 is a diagram showing a wheel speed frequency profile detected bya wheel speed sensor in certain driving conditions.

FIG. 14 is a control block diagram showing a configuration of roll rateminimization control in the first embodiment.

FIG. 15 is a time chart showing an envelope waveform shaping process forroll rate minimization control in the first embodiment.

FIG. 16 is a block diagram showing a control configuration for unsprungmass vibration damping control in the first embodiment.

FIG. 17 is a control block diagram showing a control configuration for adamping force control unit of the first embodiment.

FIG. 18 is a flow chart of a damping coefficient reconciliation processperformed during a standard mode in the first embodiment.

FIG. 19 is a flow chart of a damping coefficient reconciliation processperformed during a sport mode in the first embodiment.

FIG. 20 is a flow chart of a damping coefficient reconciliation processperformed during a comfort mode in the first embodiment.

FIG. 21 is a flow chart of a damping coefficient reconciliation processperformed during a highway mode in the first embodiment.

FIG. 22 is a time chart showing changes in damping coefficient whendriving on a hilly road surface and a bumpy road surface.

FIG. 23 is a flow chart of a driving state-based mode selection processperformed by a damping coefficient-reconciling unit of the firstembodiment.

FIG. 24 is a control block diagram showing a control configuration ofthe vehicle control device according to the first embodiment.

FIG. 25 is a control block diagram showing a configuration of adeterioration estimation accuracy detected control unit in the firstembodiment.

FIG. 26 is a descriptive diagram showing a method of setting adeterioration estimation accuracy damping coefficient, in the dampingcoefficient setting unit of the first embodiment.

DETAILED DESCRIPTION OF THE EMBODIMENTS First Embodiment

FIG. 1 is a schematic system diagram showing a vehicle control deviceaccording to a first embodiment. A vehicle comprises an engine 1constituting a power source, brakes 20 for generating braking torque byapplying frictional force to the wheels (brakes corresponding toindividual wheels will be referred to hereafter as follows: front rightbrake: 20FR; front left brake: 20FL; rear right brake: 20RR; rear leftbrake: 20RL), and shock absorbers 3 capable of variable damping forcecontrol, provided between each of the wheels and the vehicle body(“shock absorber” will be abbreviated “S/A” in the followingdescription, and S/A corresponding to individual wheels will be referredto as follows: front right S/A: 3FR; front left S/A: 3FL; rear rightS/A: 3RR; rear left S/A: 3RL).

The engine 1 has an engine controller (also referred to hereinafter asan engine control unit, and corresponding to the power source controlmeans) 1 a. The engine controller la controls the engine operation state(engine rpm, engine output torque, etc.) as desired by controlling theopening of the throttle valve, the fuel injection level, the ignitiontiming, and the like, of the engine 1. The brakes 20 generate brakingtorque based on hydraulic pressure supplied from a brake control unit 2capable of controlling brake hydraulic pressure for each of the wheels,according to the driving state. The brake control unit 2 has a brakecontroller 2 a (also referred to hereinafter as a brake control unit)for controlling the braking torque generated by the brakes 20, thedesired hydraulic pressure being generated in the brakes 20 for each ofthe wheels through opening and closing of a plurality of solenoid valvesusing master cylinder pressure generated by a driver operating the brakepedal, or pump pressure generated by a built-in motor-driven pump, as ahydraulic pressure source.

The S/A 3 are damping force-generating devices for damping the elasticmotion of coil springs provided between the unsprung mass (the axles,wheels, etc.) and the sprung mass (vehicle body, etc.) of the vehicle,the damping force being designed to be variable through the operation ofactuators. Each S/A 3 has a cylinder in which fluid is sealed, a pistonthat makes strokes within the cylinder, and an orifice for controllingthe movement of the fluid between fluid chambers formed above and belowthe piston. Orifices having different orifice diameters are formed inthe piston, and an orifice corresponding to a control command isselected from among the various orifices when the S/A actuator operates.Damping force corresponding to the diameter of the orifice is therebygenerated. The movement of the piston will be more easily restricted ifthe orifice diameter is small, whereby the damping force will be higher,and movement of the piston will be less easily restricted if the orificediameter is large, decreasing the damping force.

Apart from selecting the diameter of the orifice, damping force may alsobe set, for example, by disposing a solenoid control valve on acommunicating passage connecting the fluid, formed above and below thepiston, and controlling the amount of opening and closing of thesolenoid control valve; the invention is not particularly limited withrespect thereto. Each S/A 3 has an S/A controller 3 a (corresponding toa damping force control means) for controlling the damping force of theS/A 3, the damping force being controlled through operation of theorifice diameter by the S/A actuator.

Also provided are wheel speed sensors 5 for detecting the wheel speed ofeach of the wheels (hereinbelow, when indicating wheel speedscorresponding to individual wheels, these will be referred to as: frontright wheel speed: 5FR; front left wheel speed 5FL; rear right wheelspeed: SRR; rear left wheel speed: 5RL); an integrated sensor 6 fordetecting forward/reverse acceleration, yaw rate, and lateralacceleration acting upon the center of gravity of the vehicle; asteering angle sensor 7 for detecting a steering angle indicating theamount to which the driver has operated the steering wheel; a vehiclespeed sensor 8 for detecting vehicle speed, an engine torque sensor 9for detecting engine torque; an engine rpm sensor 10 for detectingengine rpm; a master pressure sensor 11 for detecting master cylinderpressure; a brake switch 12 for outputting an on state signal when abrake pedal is operated; an accelerator opening sensor 13 for detectingthe opening degree of the accelerator pedal; and a temperature sensor 14for detecting outside air temperature. Signals from these varioussensors are inputted to the engine controller 1, the brake controller 2a, and the S/A controller 3 a, as needed. The integrated sensor 6 may bedisposed at the location of the center of gravity of the vehicle, or atsome other location, with no particular limitation, provided that theconfiguration is on by which estimation of various values at theposition of the center of gravity is possible. The sensor need not be ofintegrated type; individual sensors for detecting yaw rate,forward/reverse acceleration, and lateral acceleration may also beprovided.

(Overall Configuration of Vehicle Control Device)

In the vehicle control device according to the first embodiment, threeactuators are used in order to control the vibrational state arising inthe sprung mass. Because control respectively performed thereby servesto control the state of the sprung mass at this time, interference is aproblem. In addition, the elements controllable by the engine 1, theelements controllable by the brakes 20, and the elements controllable bythe S/A 3 respectively differ, and the issue which combinations theseelements should be controlled in is another problem. For example, thebrakes 20 are capable of controlling bouncing motion and pitchingmotion, but controlling both will create a strong sense of decelerationand tend to create discomfort for the driver. The S/A 3 are capable ofcontrolling all rolling motion, bouncing motion, and pitching motion,but in cases in which all over wide ranges is performed by the S/A 3,the S/A 3 manufacturing costs may be greater, and the damping force willtend to be higher, creating a tendency for high-frequency vibration tobe input from the road surface, and tending to produce discomfort forthe driver. In other words, there is a trade-off in that controlperformed by the brakes 20 will not lead to a worsening ofhigh-frequency vibration, but will lead to an increased sense ofdeceleration, whereas control performed by the S/A 3 will not create asense of deceleration, but may lead to input of high-frequencyvibration.

Thus, in the vehicle control device of the first embodiment, acomprehensive assessment is made of these problems, and in order toachieve a control configuration utilizing the respective advantages ofthe respective control characteristics while compensating for theweaknesses of each other, to thereby achieve a vehicle control devicethat, while inexpensive, offers superior vibration damping performance,the control system as a whole was built in consideration primarily ofthe following exemplary points.

(1) The amount of control by the S/A 3 is minimized, through parallelcontrol of the engine 1 and the brakes 20.

(2) The motion subjected to control by the brakes 20 is limited topitching motion, thereby eliminating the sense of deceleration producedby control by the brakes 20.

(3) The amount of control performed by the engine 1 and the brakes 20 islimited to less than the actually outputtable control amount, therebyreducing the load on the S/A 3, while minimizing the unnatural feelassociated with control performed by the engine 1 and the brakes 20.

(4) Skyhook control is performed by all of the actuators. This allowsskyhook control to be accomplished through an inexpensive configurationutilizing all of the wheel speed sensors installed in the vehicle,without the use of a stroke sensor, a sprung mass vertical accelerationsensor, or the like, as is usually necessary for skyhook control.

(5) Scalar control (frequency-sensitive control) has been newlyintroduced in order to address input of high-frequency vibration, whichis difficult to address using skyhook control or other types of vectorcontrol, during sprung mass control by the S/A 3.

(6) The control state to be accomplished by the S/A 3 is selected, asappropriate, according to the driving state, thereby providing a controlstate suited to the driving conditions.

The foregoing summarizes the features of the control system in itsentirety as constituted according to the embodiment. The specifics bywhich these individual features will be described in sequence hereafter.

FIG. 2 is a control block diagram showing a configuration of control bythe vehicle control device according to the first embodiment. In thefirst embodiment, the control devices are constituted by three units: anengine controller 1 a, a brake controller 2 a, and an S/A controller 3a, with a wheel speed feedback control system being configured in therespective control devices. Separately from the respective controldevices, the system has an estimation accuracy deterioration detectionunit 4 a for detecting deterioration in the estimation accuracy, whichindicates the reliability of the state estimations made by the drivingstate estimating units, discussed later (a first driving stateestimating unit 100, a second driving state estimating unit 200, and athird driving state estimating unit 32); and a deterioration estimationaccuracy detected control unit 5 a, for transitioning to a suitablecontrol state when deterioration in the estimation accuracy has beendetected. The estimation accuracy deterioration detection unit 4 a andthe deterioration estimation accuracy detected control unit 5 a will bedescribed in detail below. Here, in the first embodiment, aconfiguration provided with three control devices as the control devicesis shown; however, a configuration in which all of the control devicesare constituted by an integrated control device would be acceptable aswell, with no particular limitations. The configuration provided withthree control devices in the first embodiment envisions repurposing theengine controller and the brake controller of an existing vehicle as theengine control unit 1 a and the brake control unit 2 a, while installingthe separate S/A controller 3 a, to realize the vehicle control deviceaccording to the first embodiment.

(Configuration of Engine Controller)

The engine controller 1 a has the first driving state estimating unit100 for estimating, based on wheel speed detected primarily by wheelspeed sensors, the stroke speed, bounce rate, roll rate, and pitch rateof each wheel, for use in skyhook control by a sprung mass vibrationdamping control unit 101 a, discussed below; an engine orientationcontrol unit 101 for calculating an engine orientation control amountconstituting an engine torque command; and an engine control unit 102for controlling the operation state of the engine 1, based on thecalculated engine orientation control amount. The specifics of theestimation process by the first driving state estimating unit 100 willbe discussed below. The engine orientation control unit 101 has thesprung mass vibration damping control unit 101 a which calculates asprung mass control amount for minimizing bouncing motion and pitchingmotion caused by skyhook control; a ground-contacting load control unit101 b that calculates a ground-contacting load fluctuation-minimizingcontrol amount for minimizing ground-contacting load fluctuation of thefront wheels and the rear wheels; and an engine-side driver inputcontrol unit 101 c for calculating, based on signals from the steeringangle sensor 7 and the vehicle speed sensor 8, a yaw response controlamount corresponding to the vehicle behavior that the driver wishes toachieve. The engine orientation control unit 101 calculates an engineorientation control amount at which the control amounts calculated bythe control units reach minimum, doing so through optimal control (LQR),and outputs a final engine orientation control amount to the enginecontrol unit 102. By minimizing bouncing motion and pitch motion by theengine 1 in this manner, the amount of damping force control in the S/A3 can be reduced, and exacerbation of high-frequency vibration can beavoided. Because the S/A 3 can focus on minimizing rolling motion,rolling motion can be effectively minimized.

(Configuration of Brake Controller)

The brake controller 2 a has the second driving state estimating unit200, which based on wheel speed detected by the wheel speed sensors 5,estimates a stroke speed, a pitch rate, and so on for each wheel; askyhook control unit 201 which, based on skyhook control based on theestimated stroke speed and pitch rate, calculates a brake orientationcontrol amount (discussed in detail later); and a brake control unit 202that controls braking torque of the brakes 20, based on the calculatedbrake orientation control amount. In the first embodiment, identicalprocesses are employed as the estimation processes in the first drivingstate estimating unit 100 and the second driving state estimating unit200; however, other estimation processes could be employed, providedthat the processes involve estimation from wheel speed. By havingpitching motion be controlled by the brakes 20 in this manner, theamount of damping force control in the S/A 3 can be reduced, andexacerbation of high-frequency vibration can be avoided. Because the S/A3 can focus on minimizing rolling motion, rolling motion can beeffectively minimized.

(Configuration of S/A Control Device)

The S/A controller 3 a has a driver input control unit 31 that performsdriver input control to reach a desired vehicle orientation based on anoperation by the driver (a steering operation, accelerator operation,brake pedal operation, or the like); a third driving state estimatingunit 32 that estimates a driving state based on values detected byvarious sensors (primarily the wheel speed sensor values of the wheelspeed sensors 5); a sprung mass vibration damping control unit 33 thatcontrols the vibration state of the sprung mass based on the estimateddriving state; an unsprung mass vibration damping control unit 34 thatcontrols the vibration state of the unsprung mass based on the estimateddriving state; and a damping force control unit 35 that, based on ashock absorber orientation control amount output by the driver inputcontrol unit 31, a sprung mass vibration damping amount output by thesprung mass vibration damping control unit 33, and an unsprung massvibration damping amount output by the unsprung mass vibration dampingcontrol unit 34, determines damping force to be set for the S/A 3, andperforms damping force control of the S/A. In the first embodiment,identical estimation processes are employed as the estimation processesin the first driving state estimating unit 100, the second driving stateestimating unit 200, and the third driving state estimating unit 32;however, other estimation processes could be employed, with noparticular limitations, provided that the processes involve estimationfrom wheel speed.

In the first embodiment, the feedback control mechanisms are constitutedto employ the wheel speed sensors 5 of all of the actuators. FIG. 3 is aconceptual diagram showing a configuration of a wheel speed feedbackcontrol system according to the first embodiment. The engine 1, thebrakes 20, and the S/A 3 respectively constitute a separate enginefeedback control system, brake feedback control system, and SA feedbackcontrol system. When the respective actuators are operated separatelywith no monitoring of their reciprocal operation states, controlinterference becomes a problem. However, because the effects produced bycontrol of each actuator are manifested as fluctuations of respectivewheel speeds, by constituting a wheel speed feedback control system, theeffects of the actuators are reciprocally monitored as a result,avoiding control interference. For example, when certain sprung massvibration is minimized by the engine 1, fluctuations in wheel speedarise in association therewith. Even where the other actuators are notaware of the specifics of control performed by the engine 1, control bythe brakes 20 and the S/A 3 is performed based on wheel speed, whichreflects these effects. Specifically, because the feedback controlmechanism is constituted using the shared values of wheel speed, evenduring separate control without the application of reciprocal monitoringfor control purposes, control is ultimately performed based onreciprocal monitoring (hereinafter, such control is termed cooperativecontrol), and the vehicle orientation can be made to converge in astabilized direction. The feedback control systems will be described insuccession below.

(Driving State Estimating Unit)

First, the first, second, and third driving state estimating units whichare provided to the feedback control systems and share a commonconfiguration, will be described. In the first embodiment, identicalestimation processes are employed as the estimation processes in thefirst driving state estimating unit 100, the second driving stateestimating unit 200, and the third driving state estimating unit 32.Therefore, because the processes in the estimating units are common toeach, the estimation process taking place in the third driving stateestimating unit 32 will be described as representative. Provided thatwheel speed is employed in state estimation, each of these driving stateestimating units may be provided with a separate estimation model, withno particular limitations.

FIG. 4 is a control block diagram showing the configuration of the thirddriving state estimating unit of the first embodiment. The third drivingstate estimating unit 32 of the first embodiment calculates a strokespeed, bounce rate, roll rate, and pitch rate for each wheel used inskyhook control performed by the sprung mass vibration damping controlunit 33 as described below, doing so fundamentally based on the wheelspeeds detected by the wheel speed sensors 5. First, the values from thewheel speed sensor 5 of each of the wheels are inputted into a strokespeed calculator unit 321, and sprung mass speed is calculated by thestroke speed calculator unit 321 from the stroke speeds calculated forthe wheels.

FIG. 5 is a control block diagram showing the specifics of control in astroke speed calculator unit of the first embodiment. A stroke speedcalculator unit 321 is separately provided for each wheel; the controlblock diagram shown in FIG. 5 is a control block diagram focusing on aspecific wheel. The stroke speed calculator unit 321 has a referencewheel speed calculating unit 300 for calculating a reference wheel speedbased on the values from the wheel speed sensors 5, a front wheelsteering angle of detected by the steering angle sensor 7, a rear wheelsteering angle δr (the actual rear wheel steering angle when a rearwheel steering device is provided; otherwise zero), a vehicle bodylateral speed, and an actual yaw rate detected by the integrated sensor6; a tire rotational vibration frequency calculating unit 321 a forcalculating tire rotational vibration frequency based on the calculatedreference wheel speed; a deviation calculating unit 321 b forcalculating the deviation between the reference wheel speed and wheelspeed sensor values (i.e., wheel speed variation); a GEO conversion unit321 c for converting the deviation calculated by the deviationcalculating unit 321 b to a suspension stroke amount; a stroke speedcalibrating unit 321 d for calibrating the converted stroke amount to astroke speed; and a signal processing unit 321 e for applying a bandelimination filter corresponding to the frequency calculated by the tirerotational vibration frequency calculating unit 321 a to the calibratedvalue from the stroke speed calibrating unit 321 d, to eliminate aprimary tire rotational vibration component, and calculate a finalstroke speed.

(Reference Wheel Speed Calculator Unit)

The reference wheel speed calculator unit 300 will now be described.FIG. 6 is a block diagram showing the configuration of a reference wheelspeed calculator unit of the first embodiment. The reference wheel speedindicates a value from which various types of interference have beeneliminated from each of the wheel speeds. In other words, the differencebetween a wheel speed sensor value and the reference wheel speed is avalue related to a component that varies according to a stroke generatedby vehicle body bouncing behavior, rolling behavior, pitching behavior,or unsprung vertical vibration; in the present embodiment, the strokespeed is calculated based on this difference.

A flat surface motion component extractor unit 301, using the wheelspeed sensor values as inputs, calculates a first wheel speed V0 as areference wheel speed for each of the wheels based on a vehicle bodyplan view model. ω (rad/s) is the wheel speed sensor value detected bythe wheel speed sensor 5, δf (rad) is a front wheel actual steeringangle detected by the steering angle sensor 7, δr (rad) is a rear wheelactual steering angle, Vx is vehicle body lateral speed, γ (rad/s) isthe yaw rate detected by the integrated sensor 6, V (m/s) is a vehiclebody speed estimated from the calculated reference wheel speed ω0, VFL,VFR, VRL, and VRR are the reference wheel speeds to be calculated, Tf isa front wheel tread, Tr is a rear wheel tread, Lf is the distance to thefront wheels from the location of the vehicle center of gravity, and Lris the distance to the rear wheels from the location of the vehiclecenter of gravity. The vehicle body plan view model is expressed asfollows, using the symbols described above.

VFL=(V−Tf/2·γ)cos δf+(Vx+Lf·γ)sin δf

VFR=(V+Tf/2·γ)cos δf+(Vx+Lf·γ)sin δf

VRL=(V−Tr/2·γ)cos δr+(Vx−Lr·γ)sin δr

VRR=(V+Tr/2·γ)cos δr+(Vx−Lr·γ)sin δr  (Formula 1)

Assuming normal driving in which no lateral sliding of the vehicle, a“0” may be input for the vehicle body lateral speed Vx. When rewrittenwith values based on V in the respective formulas, the expressions areas follows. When rewritten in this manner, V is denoted as V0FL, V0FR,V0RL, and V0RR (equivalent to first wheel speeds) as valuescorresponding to the respective wheels.

V0FL={VFL−Lf·γ sin δf}/cos δf+Tf/2·γ

V0FR={VFR−Lf·γ sin δf}/cos δf−Tf/2·γ

V0RL={VRL+Lr·γ sin δr}/cos δR+Tf/2·γ

V0RR={VRR+Lr·γ sin δr}/cos δR−Tf/2·γ  (Formula 2)

A roll interference elimination unit 302, using the first wheel speed V0as input, calculates second wheel speeds V0F, V0R as reference wheelspeeds for the front and rear wheels based on a vehicle body front viewmodel. The vehicle body front view model is used to eliminate wheelspeed differences produced by rolling motion occurring around a centerof roll rotation on a vertical line passing through the vehicle centerof gravity, when the vehicle is viewed from the front, and isrepresented by the following formulas.

V0F=(V0FL+V0FR)/2

V0R=(V0RL+V0RR)/2

This yields second wheel speeds V0F, V0R from which roll-basedinterference has been eliminated.

A pitch interference elimination unit 303, using the second wheel speedsV0F, V0R as inputs, calculates third wheel speeds VbFL, VbFR, VbRL, andVbRR constituting reference wheel speeds for all the wheels based on avehicle body side view model. Here, the vehicle body side view model isone used to eliminate wheel speed differences produced by pitchingmotion occurring around a center of pitch rotation on a vertical linepassing through the vehicle center of gravity, when the vehicle isviewed from the front, and is represented by the following formula.

VbFL=VbFR=VbRL=VbRR 32 {Lr/(Lf+Lr)}V0F+{Lf/(Lf+Lr)}V0R  (Formula 3)

A reference wheel speed redistribution unit 304 respectively substitutesVbFL (=VbFR=VbRL=VbRR) for V in the vehicle body plan view model shownin formula 1, to calculate final reference wheel speeds VFL, VFR, VRL,VRR for each wheel, each of which is divided by the tire radius r0 tocalculate a reference wheel speed ω0.

Once the reference wheel speed ω0 has been calculated for each wheelthrough the process described above, deviation between this referencewheel speed ω0 and the wheel speed sensor values is calculated; becausethis deviation represents wheel speed variation associated with thesuspension stroke, it is converted into a stroke speed Vzs. Basically,the stroke of suspension, when holding the wheels, takes place not justin the vertical direction; additionally the wheel rotational centersmove forward or backward in association with the stroke, and the axlesequipped with the wheel speed sensors 5 themselves become tilted,creating a difference in rotational angle from the wheels. Because thisforward and backward motion leads to changes in wheel speed, deviationsbetween the reference wheel speed and the wheel speed sensor values canbe extracted by way of stroke-associated fluctuations. The extent offluctuation that occurs can be set, as appropriate, according to thesuspension geometry.

Once the stroke speeds Vz_sFL, Vz_sFR, Vz_sRL, Vz_sRR for each wheelhave been calculated by the stroke speed calculator unit 321 accordingto the process described above, a sprung mass speed calculating unit 322calculates a bounce rate, a roll rate, and a pitch rate, for use inskyhook control.

(Estimation Model)

Skyhook control refers to a process whereby damping force is set basedon the relationship between the stroke speed of the S/A 3 and the sprungmass speed, and the orientation of the sprung mass is controlled toachieve a flat driving state. Here, in order to achieve control of theorientation of the sprung mass through skyhook control, feedback of thesprung mass speed is necessary. Stroke speed is a value detectable fromthe wheel speed sensor 5, and since the sprung mass is not provided witha vertical acceleration sensor or the like, the sprung mass speed mustbe estimated using an estimation model. Issues pertaining to theestimation model, and the appropriate model configuration to adopt, willbe discussed below.

FIG. 7 is a schematic diagram showing a vehicle body vibration model.FIG. 7( a) is a model for a vehicle provided with S/A having a constantdamping force (hereafter referred to as a conventional vehicle), andFIG. 7( b) is a model of a case in which variable-damping force S/A areprovided, and skyhook control is performed. In FIG. 7, Ms indicates thesprung mass, Mu indicates the unsprung mass, Ks indicates the coilspring modulus of elasticity, Cs indicates the S/A damping coefficient,Ku indicates the unsprung (tire) modulus of elasticity, Cu indicates theunsprung (tire) damping coefficient, and Cv indicates a variable dampingcoefficient. z2 indicates the position of the sprung mass, z1 indicatesthe position of the unsprung mass, and z0 indicates the position of theroad surface.

In the case of using the conventional vehicle model shown in FIG. 7( a),the equation of motion for the sprung mass is expressed as follows. Thefirst-order differential for z1 (i.e., speed) is represented by dz1, andthe second-order differential (i.e., acceleration) is represented byddzl.

Ms·ddz2=−Ks(z2−z1)−Cs(dz2−dz1)  (Estimation formula 1)

Applying a Laplace transform to this relational expression yields thefollowing formula.

dz2=−(1/Ms)·(1/s ²)·(Cs·s+Ks)(dz2−dz1)  (Estimation formula 2)

Here, dz2−dz1 is the stroke speed (Vz_sFL, Vz_sFR, Vz_sRL, Vz_sRR), andtherefore the sprung mass speed can be calculated from the stroke speed.However, modifying the damping force through skyhook control willgreatly reduce estimation precision, creating the problem that, in theconventional vehicle model, a large orientation control force (dampingforce modification) cannot be applied.

However, the use of a skyhook control-based vehicle model like thatshown in FIG. 7( b) is conceivable. Modification of the damping forcebasically involves modifying the force which limits the piston movementspeed of the S/A 3 associated with the suspension stroke. Due to thefact that semi-active S/A 3, in which the pistons cannot be activelymoved in a desired direction, are employed, a semi-active skyhook modelis adopted, and the calculation of the sprung mass speed is as follows.

dz2=−(1/Ms)·(1/s ²)·{(CS+CV)·s+Ks}(dz2−dz1)  (Estimation formula 3)

Here,

if dz2·(dz2−dz1)≧0, then Cv=Csky·{dz21(dz2−dz1)}, and

if dz2·(dz2−dz1)<0, then Cv=0.

That is, Cv is a discontinuous value.

In cases in which it is desired to estimate sprung mass speed using asimple filter in a semi-active skyhook model, when the model itself isviewed as the filter, the variables will be equivalent to the filtercoefficients, and the pseudo-differential term {(Cs+Cv)·s+Ks} includes adiscontinuous variable damping coefficient Cv; therefore, filterresponse is unstable, and suitable estimation precision cannot beobtained. In particular, unstable filter response will lead to shiftingof phase. Skyhook control cannot be accomplished if the correspondencerelationship of the phase and the sign of sprung mass speed breaks down.It was therefore decided to estimate the sprung mass speed by using anactive skyhook model, in which it is possible to use a stable Cskydirectly without relying upon the sign relationship of sprung mass speedand stroke speed, even in cases in which semi-active S/A 3 are used.Where the active skyhook model is adopted, the calculation of sprungmass speed is as follows.

dz2=−(1/s)·1/(s+Csky/Ms)}·{(Cs/Ms)s+(Ks/Ms)}(dz2−dz1)  (Estimationformula 4)

In this case, the pseudo-differential term {(Cs/Ms)s+(Ks/Ms)} does notgive rise to discontinuity, and the term {1/(s+Csky/Ms)} can beconstituted by a low-pass filter. Filter response is therefore stable,and suitable estimation precision can be obtained. It should be notedthat, despite adopting the active skyhook model here, in actual practiceonly semi-active control is possible, and thus, the controllable rangeis halved. The magnitude of the estimated sprung mass speed is thereforeless than the actual speed in the frequency band below sprung massresonance; however, as phase is the most important element in skyhookcontrol, skyhook control can be achieved as long as the correspondencebetween phase and sign can be maintained; the magnitude of the sprungmass speed can be adjusted using the other coefficients or the like, andtherefore does not pose a problem.

From the above relationship, it will be appreciated that sprung massspeed can be estimated, provided that the stroke speed of each wheel isknown. Because an actual vehicle has not one wheel, but four, estimationof the state of the sprung mass using the stroke speed of each wheel,through modal decomposition to roll rate, pitch rate, bounce rate, willbe examined next. When calculating the abovementioned three componentsfrom the stroke speeds of the four wheels, one corresponding componentis lacking, leading to an indefinite solution; thus, warp rate, whichindicates movement of diagonally opposed wheels, has been introduced.Where xsB denotes the bounce term, xsR the roll term, xsP the pitchterm, and xsW the warp term of the stroke amount, and z_sFL, z_sFR,z_sRL, z_sRR denote stroke amounts corresponding to Vz_sFL, Vz_sFR,Vz_sRL, Vz_sRR, the following formula is true.

$\begin{matrix}{\begin{Bmatrix}{z\_ sFL} \\{z\_ sFR} \\{z\_ sRL} \\{z\_ sRR}\end{Bmatrix} = {{\begin{bmatrix}1 & 1 & {- 1} & {- 1} \\1 & {- 1} & {- 1} & 1 \\1 & 1 & 1 & 1 \\1 & {- 1} & 1 & {- 1}\end{bmatrix}\begin{Bmatrix}{xsB} \\{xsR} \\{xsP} \\{xsW}\end{Bmatrix}\begin{Bmatrix}{xsB} \\{xsR} \\{xsP} \\{xsW}\end{Bmatrix}} = {\begin{bmatrix}1 & 1 & {- 1} & {- 1} \\1 & {- 1} & {- 1} & 1 \\1 & 1 & 1 & 1 \\1 & {- 1} & 1 & {- 1}\end{bmatrix}^{- 1}\begin{Bmatrix}{z\_ sFL} \\{z\_ sFR} \\{z\_ sRL} \\{z\_ sRR}\end{Bmatrix}}}} & \left( {{Formula}\mspace{14mu} 1} \right)\end{matrix}$

From the relational expression shown above, the differentials dxsB, . .. of xsB, xsR, xsP, xsW may be expressed by the following formulas.

dxsB=¼(Vz _(—) sFL+Vz _(—) sFR+Vz _(—) sRL+Vz _(—) sRR)

dxsR=¼(Vz _(—) sFL−Vz _(—) sFR+Vz _(—) sRL−VzsRR)

dxsP=¼(−Vz _(—) sFL−Vz _(—) sFR+Vz _(—) sRL+Vz _(—) sRR)

dxsW=¼(−Vz _(—) sFL+Vz _(—) sFR+Vz _(—) sRL−Vz _(—) sRR)

Here, the relationship of sprung mass speed and stroke speed is obtainedfrom estimation formula 4 above; thus, in estimation formula 4, wherethe part −(1/s)·{1/(s+Csky/Ms)}·{(Cs/Ms)s+(Ks/Ms)} is denoted as G;values which take into account modal parameters according to the bounceterm, roll term, and pitch term of Csky, Cs, and Ks, respectively(CskyB, CskyR, CskyP, CsB, CsR, CsP, KsB, KsR, KsP) are denoted as GB,GR, and GP; and the bounce rate is denoted as dB, the roll rate as dR,and the pitch rate as dP, the following values of dB, dR, and dP can becalculated.

dB=GB·dxsB

dR=GR·dxsR

dP=GP·dxsP

As shown above, estimates of the state of the sprung mass of an actualvehicle can be accomplished based on the stroke speeds for the variouswheels.

(Sprung Mass Vibration Damping Control Unit)

Next, the configuration of the skyhook control executed in the sprungmass vibration damping control unit 101 a, the skyhook control unit 201,and the sprung mass vibration damping control unit 33 will be described.During skyhook control, control is carried out to bring the sprung massstate that was estimated based on wheel speed in the aforedescribedmanner, to a target sprung mass state. In other words, changes in wheelspeed are changes corresponding to the sprung mass state, and in casesin which the sprung mass state, i.e., bounce, roll, and pitch, iscontrolled to a target sprung mass state, control is carried out suchthat changes in the detected wheel speed are equal to wheel speedchanges corresponding to the target sprung mass state.

(Configuration of Skyhook Control Unit)

The vehicle control device according to the first embodiment is providedwith three actuators for achieving sprung mass orientation control,namely, the engine 1, the brakes 20, and the S/A 3. Of these, the twoelements of bounce rate and pitch rate are targeted for control by thesprung mass vibration damping control unit 101 a in the enginecontroller 1 a; the pitch rate is targeted for control by the skyhookcontrol unit 201 in the brake controller 2 a; and the three elements ofbounce rate, roll rate, and pitch rate are targeted for control by theskyhook control unit 33 a in the S/A 3.

The skyhook control amount in the bounce direction is:

FB=CskyB·dB

The skyhook control amount in the roll direction is:

FR=CskyR·dR

The skyhook control amount in the pitch direction is:

FP=CskyP·dP

(Skyhook Control Amount FB in the Bounce Direction)

The skyhook control amount FB in the bounce direction is calculated aspart of the engine orientation control amount in the sprung massvibration damping control unit 101 a, and is also calculated as part ofthe S/A orientation control amount in the skyhook control unit 33 a.

(Skyhook Control Amount FR in the Roll Direction)

The skyhook control amount FR in the roll direction is calculated aspart of the S/A orientation control amount in the skyhook control unit33 a.

(Skyhook Control Amount FP in the Pitch Direction)

The skyhook control amount FP in the pitch direction is calculated aspart of the engine orientation control amount in the sprung massvibration damping control unit 101 a, calculated as part of the brakeorientation control amount in the skyhook control unit 201, and alsocalculated as part of the S/A orientation control amount in the skyhookcontrol unit 33 a.

In the engine orientation control unit 101, there is established a limitvalue for limiting the engine torque control amount according to theengine orientation control amount, so as to avoid discomfort for thedriver. In so doing, the engine torque control amount, when converted toforward/reverse acceleration, is limited to within a prescribedforward/reverse acceleration range. Therefore, when calculating anengine orientation control amount (engine torque control amount based onFB or FP, when a value at or above the limit value is calculated, anengine orientation control amount is output by way of a skyhook controlamount for bounce rate or pitch rate achievable at the limit value. Inthe engine control unit 102, an engine torque control amount iscalculated based on the engine orientation control amount correspondingto the limit value, and is output to the engine 1.

In the skyhook control unit 201, as in the engine 1, there isestablished a limit value for limiting the braking torque control amountso as to avoid discomfort for the driver (the limit value will bediscussed in detail below). In so doing, the braking torque controlamount, when converted to forward/reverse acceleration, is limited towithin a prescribed forward/reverse acceleration range (a limit valuederived from passenger discomfort, actuator life, or the like).Therefore, when calculating a brake orientation control amount based onFP, when a value at or above the limit value is calculated, a pitch rateminimization amount (hereinafter denoted as “brake orientation controlamount”) achievable at the limit value is output to the brake controlunit 202. In the brake control unit 202, a braking torque control amountis calculated based on the brake orientation control amountcorresponding to the limit value, and is output to the brakes 20.

(Brake Pitch Control)

Brake pitch control will now be described. Generally, the brakes 20 arecapable of controlling both bounce and pitch; thus, it may be consideredpreferable for them to control both. However, when bounce control isperformed by the brakes 20, braking force is generated in all fourwheels simultaneously, and even in directions of low priority ofcontrol, there is a rather strong sensation of deceleration relative tothe difficulty in producing a controlling effect, and this tends tosubject the driver to discomfort. Thus, a configuration in which pitchcontrol is performed in specialized fashion by the brakes 20 has beenadopted. FIG. 8 is a control block diagram showing brake pitch controlin the first embodiment. Defining m as the mass of the vehicle body, BFfas front wheel braking force, BFr as rear wheel braking force, Hcg asthe height between the vehicle center of gravity and the road surface, aas vehicle acceleration, Mp as pitch moment, and Vp as pitch rate, thefollowing relationships are true.

BFf+BFr=m·a

m·a·Hcg=Mp

Mp=(BFf+BFr)·Hcg

If braking force is applied when the pitch rate Vp is positive, i.e.,when the front wheel side of the vehicle is sinking, the front wheelside will sink further, exacerbating the pitching motion; thus, brakingforce is not applied in such cases. On the other hand, when the pitchrate Vp is negative, i.e., the front wheel side of the vehicle islifting, the braking pitch moment imparts braking force, minimizing liftat the front wheel side. This ensures the field of view for the driver,making it easier to see ahead, and contributing to an improvedperception of safety and of flatness of ride. From the preceding, thefollowing control amounts are applied:

when Vp>0 (front wheels sinking), Mp=0

when Vp≦0 (front wheels rising), Mp=CskyP·Vp

In so doing, braking torque is generated only when the front side of thevehicle rises, allowing the sense of deceleration produced thereby to bereduced, than when braking torque is generated both when the front sidelifts and when it sinks. In addition, the frequency of operation of theactuators need only half the usual, allowing low-cost actuators to beused.

Based on the relationships described above, a brake orientation controlamount calculating unit 334 is constituted from the following controlblocks. In a dead band process sign determining unit 3341, the sign ofthe input pitch rate Vp is determined; when the sign is positive, nocontrol is necessary, so a “0” is output to a decelerationperception-reduction process unit 3342, and when the sign is negative,control is determined to be possible, and a pitch rate signal is outputto the deceleration perception-reduction process unit 3342.

(Deceleration Sense Reduction Process)

Next, a deceleration sense reduction process will be described. Thisprocess is one that corresponds to the limit created by theaforementioned limit value that was set in the brake orientation controlamount calculating unit 334. A squaring process unit 3342 a squares thepitch rate signal. This reverses the sign, and smooths the rise incontrol force. A pitch rate square damping moment calculating unit 3342b multiplies the squared pitch rate by a skyhook gain CskyP for thepitch term, which takes into account the squaring process, andcalculates the pitch moment Mp. A target deceleration calculating unit3342 c divides the pitch moment Mp by the mass m and the height Hcgbetween the vehicle center of gravity and the road surface, andcalculates target deceleration.

A jerk threshold value limiting unit 3342 d determines whether the rateof change of the calculated target deceleration, i.e., jerk, is withinpre-established ranges for a deceleration jerk threshold value and arelease jerk threshold value, and whether the target deceleration iswithin a forward/reverse acceleration limit value range. When any of thethreshold values is exceeded, the target deceleration is corrected to avalue within the ranges for the jerk threshold values. When the targetdeceleration exceeds the limit value, it is set to within the limitvalue. It is thereby possible to produce a rate of deceleration suchthat discomfort for the driver does not result.

A target pitch moment conversion unit 3343 multiplies the limited targetdeceleration calculated by the jerk threshold value limiting unit 3342d, by the mass m and the height Hcg, and calculates a target pitchmoment, which is outputted to the brake control unit 2 a.

(Frequency-Sensitive Control Unit)

Next, a frequency-sensitive control process performed in the sprung massvibration damping control unit will be described. In the firstembodiment, basically, the sprung mass speed is estimated based on thevalues detected by the wheel speed sensors 5, and skyhook control isperformed on the basis thereof, to thereby accomplish sprung massvibration damping control. However, there are cases in which,conceivably, sufficient estimation accuracy from the wheel speed sensors5 is not guaranteed, and cases in which it is desirable to activelyguarantee a comfortable driving state (i.e., a soft ride rather than avehicle body flat sensation) depending on driving conditions or thedriver's wishes. In such cases, with vector control in which therelationship (phase, etc.) of the signs of the stroke speed and thesprung mass speed is important, such as in the case of skyhook control,owing to slight phase shifts, it may prove difficult to bring aboutsuitable control, and therefore frequency-sensitive control, involvingsprung mass vibration damping control according to vibration profilescalar quantities, is introduced.

FIG. 9 is a diagram simultaneously depicting a wheel speed frequencyprofile detected by a wheel speed sensor, and a stroke frequency profilefrom a stroke sensor, not installed in the present embodiment. Here,“frequency profile” refers to a profile in which the magnitude ofamplitude versus the frequency is plotted on the y axis as a scalarquantity. A comparison of the frequency component of the wheel speedsensor 5 and the frequency component of the stroke sensor shows thatroughly similar scalar quantities can be plotted from the sprung massresonance frequency component to the unsprung mass resonance frequencycomponent. Thus, the damping force has been set based on this frequencyprofile, from among the values detected by the wheel speed sensor 5.Here, the region in which the sprung mass resonance frequency componentlies is a frequency region in which a passenger has a perception offloating in air due to swaying of the passenger's entire body, or statedanother way, a perception that gravitational acceleration acting thepassenger has decreased, and is designated as a float region (0.5-3 Hz).A region between the sprung mass resonance frequency component and theunsprung mass resonance frequency component is a frequency region inwhich, although there is no perception of reduced gravitationalacceleration, there is a perception resembling quick, frequent bouncingexperienced by a person on horseback when riding at a trot, or statedanother way, a perception of up-and-down motion which the entire body iscapable of following, and is referred to as the bounce region (3-6 Hz).The region in which the unsprung mass resonance frequency component liesis a frequency region in which, although vertical movement to an extentthat is followed by the body's mass is not experienced, quiveringvibration is transmitted to part of the passenger's body, i.e., thethighs, and is referred to as a flutter region (6-23 Hz).

FIG. 10 is a control block diagram showing frequency-sensitive controlin sprung mass vibration damping control in the first embodiment. Fromthe wheel speed sensor values, a band elimination filter 350 cuts outnoise other than the vibration component used to perform control. Apredetermined frequency region splitting unit 351 splits the region intothe respective frequency bands of a float region, a bounce region, and aflutter region. A Hilbert transform processing unit 352 performs aHilbert transform upon the split frequency bands, converting them toscalar quantities (specifically, areas calculated using amplitude andfrequency band) based on the amplitude of the frequency. A vehiclevibrational system weighting unit 353 establishes weights for actualpropagation of vibration to the vehicle in the float region, the bounceregion, and the flutter region frequency bands. A human sensationweighting unit 354 establishes weights for propagation of vibration topassengers in the float region, the bounce region, and the flutterregion frequency bands.

The establishment of human sensation weights will now be described. FIG.11 is a correlation graph showing a human sensation profile plottedagainst frequency. As shown in FIG. 11, passenger sensitivity tofrequencies is comparatively low in the float region, which is alow-frequency region, with sensitivity gradually increasing in thecourse of transition to regions of higher frequency. Frequencies inhigh-frequency regions at and above the flutter region becomeprogressively harder to transmit to the passenger. In view of this, thefloat region human sensation weight Wf is set to 0.17, the bounce regionhuman sensation weight Wh is set to 0.34 which is higher than Wf, andthe flutter region human sensation weight Wb is set to 0.38 which ishigher than Wf and Wh. It is thereby possible to increase thecorrelation between the scalar quantities of the various frequencybands, and the vibration actually propagated to passengers. These twoweighting factors may be modified, as appropriate, according to vehicleconcept or passenger preferences.

From among the frequency band weights, the weight-determining means 355calculates the proportions occupied by the weight for each of thefrequency bands. Defining a as the float region weight, b as the bounceregion weight, and c as the flutter region weight, the weighting factorfor the float region is (a/(a+b+c)), the weighting factor for the bounceregion is (b/(a+b+c)), and the weighting factor for the flutter regionis (c/(a+b+c)). A scalar quantity calculating unit 356 multiplies thescalar quantities of the frequency bands calculated by the Hilberttransform processing unit 352, by weights calculated in theweight-determining means 355, and outputs final scalar quantities. Theprocess up to this point is performed on the wheel speed sensor valuesfor each of the wheels.

A maximum value-selection unit 357 selects the maximum value from amongthe final scalar quantities calculated for each of the four wheels. Thevalue 0.01 appearing at the bottom has been established to avoid having0 as a denominator, as the total of the maximum values is used as adenominator in a subsequent process. A proportion calculating unit 358calculates a proportion, using the total of the maximum scalar quantityvalues for each of the frequency bands as the denominator, and themaximum scalar quantity value of the frequency band corresponding to thefloat region as the numerator. In other words, the proportion ofcontamination (hereafter, simply “proportion”) in the float regioncontained in all vibration components is calculated. A sprung massresonance filter 359 performs a filter process having a sprung massresonance frequency of roughly 1.2 Hz on the calculated proportion, andextracts from the calculated proportion a sprung mass resonancefrequency band component representing the float region. In other words,because the float region exists at approximately 1.2 Hz, it is believedthat the proportion of this region will also vary around 1.2 Hz. Thefinal extracted proportion is then output to the damping force controlunit 35, and a frequency-sensitive damping force control amount inaccordance with the proportion is output.

FIG. 12 is a plot showing the relationship between damping force and theproportion of vibration contamination of the float region, produced byfrequency-sensitive control in the first embodiment. As shown in FIG.12, a high damping force level is established when the float regionoccupies a large proportion, thereby reducing the vibration level ofsprung mass resonance. Even when high damping force is established,because the proportions of the bounce region and the flutter region aresmall, no high-frequency vibration or bouncy vibration is transmitted topassengers. Meanwhile, establishing a low level of damping force whenthe float region proportion is small reduces the vibration transmissionprofile at and above the sprung mass resonance, minimizinghigh-frequency vibration and yielding a smooth ride.

The advantages of frequency-sensitive control in a comparison offrequency-sensitive control and skyhook control will now be described.FIG. 13 is a diagram showing a wheel speed frequency profile detected bya wheel speed sensor 5 under certain driving conditions. This profile isespecially observed during driving on a road surface having continuoussmall irregularities, such as cobbles. When skyhook control is performedwhile driving on a road surface exhibiting this profile, the problemarises that, because in skyhook control, the damping force is determinedfrom the peak amplitude value, any degradation in phase estimation forhigh-frequency vibrational input will cause an extremely high dampingforce to be established at incorrect timing, leading to exacerbation ofhigh-frequency vibration. By contrast, if control is performed based onscalar quantities rather than vectors, as in frequency-sensitivecontrol, the float region occupies a small proportion on road surface asshown in FIG. 13, leading to establishment of low damping force. Thus,even if the amplitude of flutter region vibration is high, the vibrationtransmission profile is sufficiently reduced, allowing the exacerbationof high-frequency vibration to be avoided. As shown by the foregoing,high-frequency vibration can be minimized through frequency-sensitivecontrol based on scalar quantities, in regions where control isdifficult due to degraded phase estimation precision even if skyhookcontrol is performed using expensive sensors.

(S/a Driver Input Control Unit)

The S/A driver input control unit will be described next. In the S/Adriver input control unit 31, based on signals from the steering anglesensor 7 and the vehicle speed sensor 8, a driver input damping forcecontrol amount, corresponding to the vehicle behavior the driver wishesto achieve, is calculated and is output to the damping force controlunit 35. For example, when the nose end of the vehicle rises duringturning by the driver, the driver's field of view will tend to divergefrom the road surface, and therefore in this case, four-wheel dampingforce is output as the driver input damping force control amount, so asto prevent the nose from rising. Additionally, a driver input dampingforce control amount for minimizing roll occurring during the turn isoutput.

(Roll Control Through S/a Side Driver Input Control)

Control to minimize roll, performed through S/A side driver inputcontrol, will be described here. FIG. 14 is a control block diagramshowing a configuration of roll rate minimization control in the firstembodiment. In a lateral acceleration estimating unit 31 b 1, lateralacceleration Yg is estimated based on the front wheel steering angle ofdetected by the steering angle sensor 7, the rear wheel steering angleδr (the actual rear wheel steering angle when a rear wheel steeringdevice is provided; otherwise zero), and the vehicle speed VSP detectedby the vehicle speed sensor 8. This lateral acceleration Yg iscalculated from the following formula, using an estimated yaw rate valueγ.

Yg=VSP·γ

The estimated yaw rate value γ is calculated from the following formula.

$\begin{Bmatrix}\beta \\\gamma\end{Bmatrix} = {N\begin{Bmatrix}\delta_{f} \\\delta_{r}\end{Bmatrix}}$ $\begin{Bmatrix}\beta \\\gamma\end{Bmatrix} = {M^{- 1}N\begin{Bmatrix}\delta_{f} \\\delta_{r}\end{Bmatrix}}$ ${Here},{M = \begin{bmatrix}m_{11} & m_{12} \\m_{21} & m_{22}\end{bmatrix}},{N = \begin{bmatrix}n_{11} & n_{12} \\n_{21} & n_{22}\end{bmatrix}}$ m₁₁ = −(Ktf ⋅ Lf − Ktv ⋅ Lv)$m_{12} = {{- \frac{1}{V}}\left( {{{Ktf} \cdot {Lf}^{2}} - {{Ktv} \cdot {Lv}^{2}}} \right)}$m₂₁ = −2(Ktf + Ktv)$m_{22} = {{{- \frac{2}{V}}\left( {{{Ktf} \cdot {Lf}} - {{Ktv} \cdot {Lv}}} \right)} - {M \cdot V}}$n₁₁ = −Ktf ⋅ Lf n₁₂ = Ktv ⋅ Lr n₂₁ = −2 ⋅ Ktf n₂₂ = −2 ⋅ Ktv

Body slip angle β Body yaw rate γ Front wheel steering angle δf Rearwheel steering angle δr Body V Front wheel CP Ktf Rear wheel CP KtvFront wheel-center of gravity distance Lf Rear wheel-center of gravitydistance Lr Body mass M

In a 90° phase-lead component creation unit 31 b 2, the estimatedlateral acceleration Yg is differentiated, and a lateral accelerationdifferentiation value dYg is output. In a 90° phase-lag componentcreation unit 31 b 3, a component F (dYg) in which the phase of thelateral acceleration differentiation value lags by 90° is output. Thecomponent F (dYg) restores the phase of the component in which thelow-frequency region was eliminated by the 90° phase-lead componentcreation unit 31 b 2 to the phase of the lateral acceleration Yg, andrepresents the DC-cut component of the lateral acceleration Yg, that is,the transient component of the lateral acceleration Yg. In a 90°phase-lag component creation unit 31 b 4, a component F (Yg) in whichthe phase of the estimated lateral acceleration Yg lags by 90° isoutput. In a gain multiplication unit 31 b 5, the lateral accelerationYg, the lateral acceleration differentiation value dYg, the lateralacceleration DC-cut component F (dYg), and the 90° phase-lag component F(Yg) are respectively multiplied by a gain. Each gain is establishedbased on a roll rate transmission coefficient for the steering angle.Moreover, each gain may be adjusted according to four control modes,discussed below. In a square calculating unit 31 b 6, the componentshaving been multiplied by gain are squared and output. In a synthesisunit 31 b 7, the values output by the square calculating unit 31 b 6 aresummed. In a gain multiplication unit 31 b 8, the squared values of thesummed components are multiplied by a gain, and output. In a square rootcalculating unit 31 b 9, the square root of the value output by the gainmultiplication unit 31 b 7 is calculated, to thereby calculate a driverinput orientation control amount for roll rate minimization controlpurposes, which is output to the damping force control unit 35.

The 90° phase-lead component creation unit 31 b 2, the 90° phase-lagcomponent creation unit 31 b 3, the 90° phase-lag component creationunit 31 b 4, the gain multiplication unit 31 b 5, the square calculatingunit 31 b 6, the synthesis unit 31 b 7, the gain multiplication unit 31b 8, and the square root calculating unit 31 b 9 correspond to a Hilberttransform unit 31 b 10 that utilizes the Hilbert transform to generatean envelope waveform.

FIG. 15 is a time chart showing an envelope waveform shaping process forroll rate minimization control in the first embodiment. When steering isinitiated by the driver at time t1, a roll rate begins to be generated.At this time, by shaping an envelope waveform through addition of the90° phase-lead component dYg, and calculating a driver input orientationcontrol amount based on a scalar quantity based on the envelopewaveform, generation of the roll rate in the initial phase of steeringcan be minimized. Further, by shaping the envelope waveform throughaddition of the lateral acceleration DC-cut component F (dYg), the rollrate generated in a transient state during initiation or completion ofsteering by the driver can be minimized efficiently. In other words, ina steady turning state in which constant roll is being generated, thedamping force is not increased excessively, and degraded ride comfort isavoided. Next, as the driver enters a steering state at time t2, the 90°phase-lead component dYg and the lateral acceleration DC-cut component F(dYg) disappear, and now the 90° phase-lag component F (Yg) is added. Atthis time, even in a case in which there is not much change in the rollrate per se in the steady turning state, subsequent to initial roll,there is generated a roll rate resonance component which corresponds toroll backlash. Supposing that the phase-lag component F (Yg) had notbeen added, the damping force would be set to a small value from time t2to time t3, posing the risk of destabilization of vehicle behavior bythe roll rate resonance component. The 90° phase-lag component F (Yg)contributes to minimizing this roll rate resonance component.

When the driver transitions from the steering state to straight forwarddriving at time t3, the lateral acceleration Yg becomes less, and theroll rate also converges on a small value. Here, because damping forcehas been solidly ensured through the action of the 90° phase-lagcomponent F (Yg), destabilization of vehicle behavior due to the rollrate resonance component can be avoided.

(Unsprung Mass Vibration Damping Control Unit)

Next, the configuration of the unsprung mass vibration damping controlunit will be described. As discussed in the context of the conventionalvehicle shown in FIG. 7( a), a resonance frequency band also exists inthe tires, as they possess both a modulus of elasticity and a dampingcoefficient. However, because a tire has a mass that is less than thatof the sprung mass, and a high modulus of elasticity as well, the bandexists to the high frequency end of the sprung mass resonance. Thisunsprung mass resonance component causes tire rumbling in the unsprungmass, potentially degrading ground contact. In addition, rumbling in theunsprung mass can be uncomfortable for passengers. Thus, damping forceis established in response to the unsprung mass resonance component, inorder to minimize unsprung mass resonance-induced rumbling.

FIG. 16 is a block diagram showing a control configuration for unsprungmass vibration damping control in the first embodiment. An unsprung massresonance component extraction unit 341 applies a band-pass filter towheel speed fluctuations output from the deviation calculating unit 321b of the driving state estimating unit 32, to extract an unsprung massresonance component. The unsprung mass resonance component is extractedfrom the region at roughly 10-20 Hz within the wheel speed frequencycomponent. An envelope waveform shaping unit 342 scalarizes theextracted unsprung mass resonance component, and using an envelopefilter shapes an envelope waveform. A gain multiplication unit 343multiplies the scalarized unsprung mass resonance component by a gain,and calculates an unsprung mass vibration damping force control amount,which is output to the damping force control unit 35. In the firstembodiment, an unsprung mass resonance component is extracted byapplying a band-pass filter to wheel speed fluctuations output from thedeviation calculating unit 321 b of the driving state estimating unit32, but it would also be acceptable to apply a band-pass filter tovalues detected by the wheel speed sensors to extract the unsprung massresonance component, or for the driving state estimating unit 32 toestimate the unsprung mass speed along with the sprung mass speed, andextract the unsprung mass resonance component.

(Configuration of Damping Force Control Unit)

Next, the configuration of the damping force control unit 35 will bedescribed. FIG. 17 is a control block diagram showing a controlconfiguration for a damping force control unit of the first embodiment.The driver input damping force control amount output from the driverinput control unit 31, the S/A orientation control amount output fromthe skyhook control unit 33 a, the frequency-sensitive damping forcecontrol amount output from the frequency-sensitive control unit 33 b,the unsprung mass vibration damping force control amount output from theunsprung mass vibration damping control unit 34, and the stroke speedcalculated by the driving state estimating unit 32 are input into anequivalent viscous damping coefficient conversion unit 35 a, whichconverts these values into an equivalent viscous damping coefficient.

Of the damping coefficients converted by the equivalent viscous dampingcoefficient conversion unit 35 a (hereafter denoted respectively as adriver input damping coefficient k1, a S/A orientation dampingcoefficient k2, a frequency-sensitive damping coefficient k3, and aunsprung mass vibration damping coefficient k4), a damping coefficientreconciling unit 35 b reconciles which damping coefficients are to beused as the basis for control, and outputs a final damping coefficient.A control signal conversion unit 35 c converts a control signal (commandcurrent value) to be sent to the S/A 3, doing so based on the strokespeed and the damping coefficient reconciled by the dampingcoefficient-reconciling unit 35 b, and outputs the signal to the S/A 3.

(Damping Coefficient-Reconciling Unit)

Next, the specifics of reconciliation performed by the dampingcoefficient reconciling unit 35 b will be described. The vehicle controldevice of the first embodiment has four control modes. The first is astandard mode which envisions a state in which a suitable turning statemay be obtained when driving on typical city streets or the like. Thesecond is a sport mode which envisions a state in which a stable turningstate may be obtained when aggressively driving along a winding road orthe like. The third is a comfort mode which envisions a state in whichpriority is given to ride comfort, such as when setting off at lowvehicle speed. The fourth is a highway mode which envisions a state ofdriving at high vehicle speeds on a freeway or the like, with numerousstraight sections.

In the standard mode, control is carried out while giving priority tounsprung mass vibration damping control by the unsprung mass vibrationdamping control unit 34, while skyhook control is performed by theskyhook control unit 33 a. In the sports mode, skyhook control by theskyhook control unit 33 a and unsprung mass vibration damping control isby the unsprung mass vibration damping control unit 34 are carried out,while giving priority to driver input control by the driver inputcontrol unit 31. In the comfort mode, control is carried out whilegiving priority to unsprung mass vibration damping control by theunsprung mass vibration damping control unit 34, whilefrequency-sensitive control is performed by the frequency-sensitivecontrol unit 33 b. In the highway mode, control is carried out whileadding a control amount for the unsprung mass vibration damping controlby the unsprung mass vibration damping control unit 34, to the skyhookcontrol performed by the skyhook control unit 33 a, while also givingpriority to the driver input control performed by the driver inputcontrol unit 31. Damping coefficient reconciliation in each of thesemodes will be described below.

(Reconciliation in Standard Mode)

FIG. 18 is a flow chart showing a damping coefficient reconciliationprocess performed in the standard mode in the first embodiment.

In step S1, a determination is made as to whether the S/A orientationdamping coefficient k2 is greater than the unsprung mass vibrationdamping coefficient k4, and, when this is the case, the process advancesto step S4, in which k2 is set as the damping coefficient.

In step S2, a scalar quantity proportion for the flutter region iscalculated based on scalar quantities of the float region, bounceregion, and flutter region, described in the context of thefrequency-sensitive control unit 33 b.

In step S3, a determination is made as to whether the proportion of theflutter region is equal to or greater than a predetermined value, whenthis is the case, the process advances to step S4, in which the lowvalue k2 is established as the damping coefficient, out of a concernthat high-frequency vibration will degrade ride comfort. On the otherhand, in cases in which the proportion of the flutter region is lessthan the predetermined value, there is no basis for concern thathigh-frequency vibration will degrade ride comfort, even if a highdamping coefficient is established, and therefore the process advancesto step S5, in which k4 is set as the coefficient.

In the standard mode, as discussed above, as a general rule, priority isgiven to unsprung mass vibration damping control, which minimizesresonance in the unsprung mass. However, when the damping force requiredfor skyhook control is less than the damping force required for unsprungmass vibration damping control, and moreover the proportion of theflutter region is large, the damping force for skyhook control isestablished so as to avoid exacerbating the high-frequency vibrationprofile associated with meeting the requirements of unsprung massvibration damping control. This allows an optimal damping profile to beobtained according to the driving state, thus avoiding high-frequencyvibration-induced degradation of ride comfort, while simultaneouslyachieving a flat vehicle body feel.

(Reconciliation in Sport Mode)

FIG. 19 is a flow chart showing a damping coefficient reconciliationprocess performed during the sport mode in the first embodiment.

In step S11, the damping force distribution factors for the four wheelsare calculated based on the driver input damping coefficients k1 for thefour wheels established through driver input control. Defining k1fr asthe front right wheel driver input damping coefficient, k1fl as thefront left wheel driver input damping coefficient, k1rr as the rearright wheel driver input damping coefficient, k1rl as the rear leftwheel driver input damping coefficient, and xfr, xfl, xrr, and xrl asthe damping force distribution factors for each of the wheels,distribution factors are calculated as follows:

xfr=k1fr/(k1fr+k1fl+k1rr+k1rl)

xfl=k1fl/(k1fr+k1fl+k1rr+k1rl)

xrr=k1rr/(k1fr+k1fl+k1rr+k1rl)

xrl=k1rl/(k1fr+k1fl+k1rr+k1rl)

In step S12, a determination is made as to whether a damping forcedistribution factor x is within a predetermined range (greater than αand less than β), and when this is the case, a determination thatdistribution substantially equal for all the wheels is made, and theprocess advances to step S13; but if even one factor is outside thepredetermined range, the process advances to step S16.

In step S13, a determination is made as to whether the unsprung massvibration damping coefficient k4 is greater than the driver inputdamping coefficient k1, and when this is the case, the process advancesto step S15, in which k4 is set as a first damping coefficient k. On theother hand, when the unsprung mass vibration damping coefficient k4 isequal to or less than the driver input damping coefficient k1, theprocess advances to step S14, in which k1 is set as the first dampingcoefficient k.

In step S16, a determination is made as to whether the unsprung massvibration damping coefficient k4 equals the maximum value max settablefor the S/A 3; and when it is the case that it equals the maximum valuemax, the process advances to step S17, or if not the case, the processadvances to step S18.

In step S17, a damping force coefficient such that the maximum value ofthe driver input damping coefficients k1 for the four wheels equals theunsprung mass vibration damping coefficient k4, and that satisfies adamping force distribution factor, is calculated as the first dampingcoefficient k. In other words, a value that increases the dampingcoefficient to maximum while satisfying the damping force distributionfactor is calculated.

In step S18, a damping coefficient that satisfies the damping forcedistribution factor within a range in which the driver input dampingcoefficient k1 for each of the four wheels is equal to or greater thank4 is calculated. In other words, a value that satisfies the dampingforce distribution factor established through driver input control, andthat satisfies requirements at unsprung mass vibration damping controlend are met, is calculated.

In step S19, a determination is made as to whether the first dampingcoefficients k set in the abovementioned steps are less than the S/Aorientation damping coefficient k2 established through skyhook control;in the case of being determined to be smaller, the damping coefficientrequired at skyhook control side is larger, and therefore the processadvances to step S20, and k2 is set. On the other hand, if k is equal toor greater than k2, k is set and the process advances to step S21, and kis set.

In the sport mode, as discussed above, as a general rule, priority isgiven to unsprung mass vibration damping control, which minimizesresonance of the unsprung mass. However, because the damping forcedistribution factor required from the driver input control side isintimately related to vehicle body orientation, and in particular isprofoundly related to change in the driver's line of sight due to theroll mode, utmost priority is given to ensuring the damping forcedistribution factor, rather than the damping coefficient required fromthe driver input control side as such. For movement that induces changein vehicle body orientation in a state in which the damping forcedistribution factor is maintained, stable vehicle body orientation canbe preserved through selection of “select high” for skyhook control.

(Reconciliation in Comfort Mode)

FIG. 20 is a flow chart showing a damping coefficient reconciliationprocess performed during the comfort mode in the first embodiment.

In step S30, a determination is made as to whether thefrequency-sensitive damping coefficient k3 is greater than the unsprungmass vibration damping coefficient k4, and, in the case of beingdetermined to be greater, the process advances to step S32 in which thefrequency-sensitive damping coefficient k3 is set. On the other hand, inthe case of a determination that the frequency-sensitive dampingcoefficient k3 equal to or less than the unsprung mass vibration dampingcoefficient k4, the process advances to step S32 in which the unsprungmass vibration damping coefficient k4 is set.

In the comfort mode, as discussed above, basically, priority is given tounsprung mass resonance damping control to minimize resonance of theunsprung mass. Because in the first place frequency-sensitive control isperformed as sprung mass vibration damping control, thereby establishingan optimal damping coefficient according to the road surface conditions,control to ensure ride comfort can be accomplished, and a sensation ofinsufficient ground contact due to rattling of the unsprung mass can beavoided through unsprung mass vibration damping control. In the comfortmode, as in the standard mode, it is acceptable for the dampingcoefficient to be switched according to the proportion of flutter in thefrequency scalar quantity. This allows for a super comfort mode in whichride comfort is better ensured.

(Reconciliation in Highway Mode)

FIG. 21 is a flow chart showing a damping coefficient reconciliationprocess performed during the highway mode in the first embodiment. Thereconciliation process from steps S11 to S18 is the same as in the sportmode, and therefore a description will be omitted.

In step S40, the S/A orientation damping coefficient k2 afforded byskyhook control is added to the reconciled first damping coefficient kafforded by the process up to step S18, and the coefficient is output.

In the highway mode, as discussed above, a value obtained by adding theS/A orientation damping coefficient k2 to the reconciled first dampingcoefficient k is used when reconciling the damping coefficient. Thisoperation will now be described with reference to the drawings. FIG. 22is a time chart showing change in the damping coefficient during drivingon a hilly road surface and a bumpy road surface. For instance, whenattempting to minimize motion giving rise to swaying movement of thevehicle body due to the effects of slight hill in the road surface whendriving at high vehicle speed, if it is attempted to achieve thisthrough skyhook control alone, it will be necessary to detect slightfluctuations in wheel speed, which requires establishing a comparativelyhigh skyhook control gain. In such cases, swaying motion can beminimized, but when there are bumps or the like on the road surface,there is a risk that the control gain will be too great, resulting inexcessively high control gain and excessive damping force control. Thisgives rise to concerns of degraded ride comfort or vehicle bodyorientation.

In contrast to this, because the first damping coefficient k is setconstantly as in highway mode, a given level of damping force can beconstantly ensured, and swaying motion of the vehicle body can beminimized, even when the damping coefficient produced through skyhookcontrol is low. Additionally, because there is no need to boost theskyhook control gain, bumps in the road surface can be dealt withappropriately through ordinary control gain. Moreover, because skyhookcontrol is performed in a state in which the damping coefficient k hasbeen established, unlike in the case of a damping coefficient limit,operation of a damping coefficient reduction step is possible in thesemi-active control region, ensuring stable vehicle orientation duringhigh-speed driving.

(Mode Selection Process)

Next, a mode selection process for selecting the aforementioned drivingmodes will be described. FIG. 23 is a flow chart showing a drivingstate-based mode selection process performed by the dampingcoefficient-reconciling unit of the first embodiment.

In step S50, based on the value from the steering angle sensor 7, adetermination is made as to whether a state of driving straight aheadexists, and if a state of driving straight ahead is determined to exist,the process advances to step S51, whereas in the case of a determinationthat a state of turning exists, the process advances to step S54.

In step S51, based on the value from the vehicle speed sensor 8, adetermination is made as to whether a predetermined vehicle speed VSP1indicating a state of high vehicle speed has been reached or exceeded,and, in the case of a determination that VSP1 has been reached orexceeded, the process advances to step S52 and standard mode isselected. On the other hand, in the case of a determination that thespeed is less than VSP1, the process advances to step S53, and comfortmode is selected.

In step S54, based on the value from the vehicle speed sensor 8, adetermination is made as to whether a predetermined vehicle speed VSP1indicating a state of high vehicle speed has been reached or exceeded,and, in the case of a determination that VSP1 has been reached orexceeded, the process advances to step S55, and highway mode isselected. On the other hand, in the case of a determination that thespeed is less than VSP1, the process advances to step S56, and sportmode is selected.

That is, standard mode is selected when driving at a high vehicle speedwhen driving straight ahead, thereby making it possible to stabilize thevehicle body orientation via skyhook control, ensure ride comfort byminimizing high-frequency vibration-induced bouncing or fluttering, andminimizing resonance in the unsprung mass. Selecting comfort mode whendriving at low speeds makes it possible to minimize resonance in theunsprung mass while minimizing the transmission of vibration such asbouncing or fluttering to passengers.

Meanwhile, highway mode is selected when driving at a high vehicle speedin a state of turning, thereby performing control using a value to whicha damping coefficient has been added; thus, high damping force isyielded as a rule. It is thus possible to minimize unsprung massresonance while actively ensuring the unsprung mass resonance duringturning via driver input control, even when traveling at a high vehiclespeed. Selecting sports mode when driving at a low vehicle speed allowsunsprung mass resonance to be minimized while actively ensuring thevehicle body orientation during turning via driver input control andperforming skyhook control as appropriate, thereby allowing for drivingwith a stable vehicle orientation.

In the first embodiment, an example of a mode selection process in whichthe driving state is detected and the mode is automatically switched hasbeen presented, but it is also possible to provide a mode switch or thelike that can be operated by a driver to select the driving mode. Thisyields ride comfort and turning performance matching the driver'sdesired driving state.

(Deterioration in Accuracy of Estimation)

Deterioration in accuracy of estimation will be described next. Asmentioned above, in the first embodiment, in each of the driving stateestimating units 100, 200, 32, a stroke speed, a bounce rate, a rollrate, and a pitch rate for use in skyhook control by the sprung massdamping control unit 101 a or 33, or by the skyhook control unit 201,are estimated for each wheel, doing so based on wheel speed detected bythe wheel speed sensors 5. However, there may be envisioned scenarios inwhich, during estimation of stroke speed or the sprung mass state fromwheel speed, the accuracy of estimation is lowered for any of variousreasons. For example, in the case of driving along a road having a low μvalue, slip is prone to occur, and it is difficult to ascertain whetheror not fluctuations in wheel speed occurring in association with thisslip are due to irregularities on the road surface. Moreover, in thecase of a road having a low μvalue, the amount of fluctuation in wheelspeed attributed to irregularities on the road surface or changes in thesprung mass state tends to be small, making it difficult todifferentiate from other types of noise. Moreover, because fluctuationsin driving/braking force also give rise to fluctuations in wheel speed,it can be difficult to distinguish between these fluctuations, andsprung mass state or stroke speed. Additionally, fluctuations in wheelspeed can arise from lateral acceleration or the yaw rate in non-linearregions, such as the friction circle limit of the tires, and can bedifficult to distinguish from other types of noise. When the accuracy ofestimation is lowered, there may be instances in which, for example, thedamping force is set to a low level despite the need to be set to a highlevel, making it difficult to stabilize the sprung mass state.

This lowering of estimation accuracy merely poses a problem in terms ofprecision, and because it does not involve any sensor malfunction oractuator malfunction, it is desirable for control to continue on withthe range it is possible to do so. Accordingly, in the first embodiment,the estimation accuracy deterioration detection unit 4 a has beenprovided to detect cases of deterioration in the estimation accuracy,and in cases of deterioration in the estimation accuracy, controlcontinues to the extent possible, while ensuring performance that is atleast as good or better than an ordinary vehicle in which vehiclevibration damping control is not performed, so as to stabilize thesprung mass behavior associated with the deterioration in the estimationaccuracy.

(Estimation Accuracy Deterioration Detection Processes)

FIG. 24 is a control block diagram showing estimation accuracydeterioration detection processes of the first embodiment. In theestimation accuracy deterioration detection unit 4 a, multipledeterioration in the estimation accuracy detection processes areexecuted based on various kinds of signals, and in a signal receptionunit 400, when deterioration in accuracy in even one of the respectiveprocesses has been detected, a deteriorated-accuracy signal is output toa deterioration in accuracy hold unit 401. In the deterioration inaccuracy hold unit 401, during the interval that thedeteriorated-accuracy signal is received, and for a predetermined timesubsequent to cutoff of the deteriorated-accuracy signal (in the case ofthe first embodiment, a one-second interval), a deterioration inaccuracy flag is set continuously to “ON.” In so doing, frequentswitching of the deterioration in accuracy flag can be minimized, whileavoiding control states based on erroneous state estimation values. Therespective estimation accuracy deterioration detection processes will bedescribed in sequence below.

(Detection Via ABS, VDC, or TCS Flags)

The vehicle of the first embodiment has an anti-skid brake control unit(hereinafter termed an ABS control unit) for detecting the state of slipof the wheels during braking, and performing pressure regulation controlto bring the slip rate to a predetermined value or below; a vehicledynamics control unit (hereinafter termed a VDC control unit) forcontrolling the brake fluid pressure of prescribed wheels, to bring aturning state of the vehicle (for example, the yaw rate) to a targetturning state; and a traction control unit (hereinafter termed a TCScontrol unit) for performing brake pressure boost control or enginetorque-down control, in order to minimize drive slip when the vehiclebegins to move, or the like.

In cases in which these control units have operated, the wheel speedfluctuations of the wheels will be affected thereby, thereby posing arisk of deterioration in the estimation accuracy. Therefore, in cases inwhich ABS flag, a VDC flag, or a TCS flag, which indicate that thesecontrols have operated, has gone “ON,” a flag-ON signal is output to abrake control flag hold unit 410. In the brake control flag hold unit410, a deteriorated-estimation accuracy signal is output during theinterval in which the flag-ON signal is received. Thedeteriorated-accuracy signal continues to be output for a predeterminedduration (in the case of the first embodiment, a five-second interval)following fall of the flag-ON signal. In so doing, a steadydeteriorated-estimation accuracy signal can be output, even in casessuch that the brake control flag repeatedly goes ON/OFF.

(Detection Based on Reference Vehicle Body Speed)

Detection based on a reference vehicle body speed shall be describednext. In the first embodiment, in the first to third driving stateestimating units 100, 200, 32, when stroke speed is estimated from wheelspeed data, a reference wheel speed is calculated in order to detect acomponent that fluctuates in association with stroke of the S/A 3. Thepurpose of doing so is to extract differentials between the referencewheel speed and the wheel speed sensor values, as a component offluctuation in association with stroke. While this reference wheel speedcan ensure precise stroke speed estimation under conditions in whichslip or the like is not occurring, when slip occurs, it becomesdifficult to distinguish whether fluctuations are those associated withstroke, or wheel speed fluctuations associated with slip. Within afrequency region including the stroke speed component, the sprung massspeed component, and the like, it is not possible to differentiate fromnoise and the like, and therefore the accuracy of the signal cannot beguaranteed. Accordingly, a low-pass filter of a frequency to thelow-frequency end (in the first embodiment, 0.5 Hz) from the vibrationfrequencies arising due to stroke speed, sprung mass speed, and the likeis applied to the reference wheel speed, and in cases in which,subsequent to application of this low-pass filter, variability of thereference wheel speed among the wheels is observed, it is detected thatwheel speed fluctuations are due to slip, and that estimation accuracyhas deteriorated.

In a reference wheel speed estimating unit 420, as described in thecontext of the reference wheel speed calculating unit of FIG. 6, a firstwheel speed V0 to serve as a reference wheel speed for each of thewheels is calculated based on the vehicle body plan view model. Here, ω(rad/s) is the wheel speed sensor detected by the wheel speed sensor 5,δf (rad) is the front wheel actual steering angle detected by thesteering angle sensor 7, δr (rad) is the rear wheel actual steeringangle, Vx is the vehicle body lateral speed, γ (rad/s) is the yaw ratedetected by the integrated sensor 6, V (m/s) is the vehicle body speedestimated from the calculated reference wheel speed ω0, VFL, VFR, VRL,and VRR are the reference wheel speeds to be calculated, Tf is the frontwheel tread, Tr is the rear wheel tread, Lf is the distance from theposition of the vehicle center of gravity to the front wheels, and Lr isthe distance from the position of the vehicle center of gravity to therear wheel. The vehicle body plan view model is expressed as follows,using the symbols described above.

VFL=(V−Tf/2·γ)cos δf+(Vx+Lf·γ)sin δf

VFR=(V+Tf/2·γ)cos δf+(Vx+Lf·γ)sin δf

VRL=(V−Tr/2·γ)cos δr+(Vx−Lr·γ)sin δr

VRR=(V+Tr/2·γ)cos δr+(Vx−Lr·γ)sin δr  (Formula 1)

Assuming normal driving in which no lateral sliding of the vehicle, a“0” may be input for the vehicle body lateral speed Vx. When rewrittenwith values based on V in the respective formulas, the expressions areas follows. When rewritten in this manner, V is denoted as V0FL, V0FR,V0RL, and V0RR (equivalent to first wheel speeds) as valuescorresponding to the respective wheels.

V0FL={VFL−Lf·γ sin δf}/cos δf+Tf/2·γ

V0FR={VFR−Lf·γ sin δf}/cos δf−Tf/2·γ

V0RL={VRL+Lr·γ sin δr}/cos δR+Tf/2·γ

V0RR={VRR+Lr·γ sin δr}/cos δR−Tf/2·γ  (Formula 2)

A reference wheel speed for each wheel is calculated based on theserelational expressions.

Next, in a low-pass filter 421, filtering at 0.5 Hz, which represents aregion to the low-frequency end from a frequency region including strokespeed and sprung mass speed is performed on the reference wheel speedV0FL, FR, RL, RR calculated for each wheel, and a stationary componentis extracted. In a differential decision unit 422, the followingrespective values are calculated.

Roll Component (Left-Right Differential)

df1=VOFL−VOFR

df1=VORL−VORR

Pitch Component (Front-Back Differential)

df3=VOFL−VORL

df4=VOFR−VORR

Warp Component (Diagonal Differential)

df5=VOFL−VPRR

df6=VOFR−VORL

Basically, in the case of calculating these differentials using valuesobtained subsequent to passage through the low-pass filter 421, as longas slip is not occurring, the reference wheel speeds of all of thewheels are equal, and therefore the differentials will be “0” orexceedingly small values. However, when slip arises, change in thestationary component arises, and therefore in cases in which anydifferential df1-df5 of a value obtained subsequent to passage throughthe low-pass filter 421 is equal to or greater than a pre-establishedprescribed value dfthi, a deteriorated-estimation accuracy signal isoutput to a reference wheel speed hold unit 423. Moreover, in order toprevent hunting during this decision, output of thedeteriorated-estimation accuracy signal is halted in cases in which thedifferential is equal to or less than value obtained by multiplying theprescribed value dfthi by 0.8. In the reference wheel speed hold unit423, for the duration of the time that the deteriorated-estimationaccuracy signal is being received, and for a prescribed durationfollowing completion of reception (in the case of the first embodiment,a two-second interval), the deteriorated-accuracy signal is outputcontinuously. In so doing, a steady deteriorated-estimation accuracysignal can be output, even in cases in which the deteriorated-estimationaccuracy signal of the differential decision unit 422 repeatedly goesON/OFF.

(Detection Based on Plan View Model)

Next, detection based on a plan view model will be described. Asdescribed previously in FIG. 14, when roll rate minimization control isperformed in driver input control, a plan view model is established, andlateral acceleration Yg is estimated.

Yg=(VSP ²/(1+A·VSP ²))·δf

Here, A is a prescribed value. The roll rate is estimated from thelateral acceleration Yg estimated based on this relationship. At thistime, under conditions such that the estimation accuracy of stroke speedis lowered due to the occurrence of slip or the like, the estimatedvalue of the aforementioned lateral acceleration will diverge from theactual value. Accordingly, a low-pass filter of a frequency (in thefirst embodiment, 0.5 to the low-frequency end from the vibrationfrequencies produced by stroke speed, sprung mass speed, or the like isapplied to the estimated lateral acceleration, and when the lateralacceleration subsequent to application of this low-pass filter divergesfrom the actual lateral acceleration detected by the lateralacceleration sensor, it is detected that wheel speed fluctuations aredue to slip, and that estimation accuracy has deteriorated.

In a vehicle motion state estimating unit 430, the vehicle speed VSPdetected by the wheel speed sensor 8 and the steering angle detected bythe steering angle sensor 7 are read in, and the lateral acceleration isestimated based on a plan view model. In like fashion, the yaw rate isestimated based on a plan view model. With regard to estimation of theyaw rate, when, for example, the yaw rate is denoted as γ, it may becalculated from the relationship Yg=VSP·γ, or estimated based on(Formula 1) or (Formula 2).

Next, in a low-pass filter 431, filtering by a low-pass filter of afrequency of 0.5 Hz, which represents a region to the low-frequency endfrom a frequency region including the stroke speed and sprung massspeed, is performed on the estimated lateral acceleration, the estimatedyaw rate, and the sensor value from the integrated sensor 6, and astationary component is extracted. Then, in a differential decision unit432, the differentials of the respective estimated value and sensorvalues are calculated.

dfyrss=estimated yaw rate−actual yaw rate

dflgss=estimated lateral acceleration−actual lateral acceleration

Basically, in cases in which these differentials dfyrss, dflgss havebeen calculated using values obtained subsequent to passage through thelow-pass filter 431, when slip or the like has not arisen, the estimatedvalues and the sensor values will be approximately equal, and the andtherefore the differentials will be “0” or exceedingly small values.However, when slip arises, change in the stationary component arises,and therefore in cases in which any differential dfyrss, dflgss of avalue obtained subsequent to passage through the low-pass filter 431 isequal to or greater than a pre-established prescribed value dfthi, adeteriorated-estimation accuracy signal is output to a plan view modelhold unit 433. Moreover, in order to prevent hunting during thisdecision, output of the deteriorated-estimation accuracy signal ishalted in cases in which the differential is equal to or less than valueobtained by multiplying the prescribed value dfthi by 0.8. In the planview model hold unit 433, for the duration of the time that thedeteriorated-estimation accuracy signal is being received, and for aprescribed duration following completion of reception (in the case ofthe first embodiment, a two-second interval), the deteriorated-accuracysignal is output continuously. In so doing, a steadydeteriorated-estimation accuracy signal can be output, even in cases inwhich the deteriorated-estimation accuracy signal of the differentialdecision unit 432 repeatedly goes ON/OFF.

(Detection Based on Shift Position)

Next, detection based on shift position will be described. In a case inwhich, for example, the reverse range has been selected, the directionof rotation of the wheels will be the opposite direction from thatduring forward advance, and change associated with wheel speedfluctuations will also differ from that during forward advance. In acase in which the parking range has been selected, because the vehicleis at a stop, there is no need to estimate the stroke speed, andestimation itself is difficult. Therefore, in a shift decision unit 440,when the shift signal indicates either the reverse range or the parkingrange, a deteriorated-accuracy signal is output continuously to a shifthold unit 441. From the standpoint of preventing hunting in the shifthold unit 441 in association with shift operations, thedeteriorated-accuracy signal is output continuously for a prescribedduration (in the first embodiment, a one-second interval) followingcompletion of reception of the deteriorated-estimation accuracy signal.

(Detection Based on Brake Switch)

Next, detection based on the brake switch will be described. When thedriver operates the brake pedal to generate braking force, followed byan operation of releasing the brake pedal, torque fluctuations occurringduring release of braking act as impulse input. Because front-backvibration is excited by this impulse input, causing the wheel speed tofluctuate, the estimation accuracy of stroke speed and sprung mass stateis lowered. Accordingly, a braking force release decision unit 450decides whether or not the brake switch has switched from ON to OFF, andin the event of a decision that it has switched, outputs adeteriorated-accuracy signal to a brake switch hold unit 451. Thedeteriorated-accuracy signal is output continuously for a prescribedduration (in the first embodiment, a one-second interval) following thepoint in time that the brake switch switched OFF.

(Detection Based on Wheel Rim Drive Torque)

Detection based on wheel rim drive torque will be described next. When asudden change in torque occurs due to sudden acceleration or shifting,changes in torque of the drive wheels, specifically, changes in thewheel rim drive torque, arise, causing the wheel speed to fluctuate.Therefore, in cases in which a change in the wheel rim drive torque,equal to or greater than a prescribed level, has occurred, it can bedecided that estimation accuracy has deteriorated. Estimation of wheelrim drive torque during acceleration can be made based on informationincluding the effective torque of the engine, the engine rpm, the rpm ofthe turbine, the rpm of the automatic transmission output shaft, and theshift position; more specifically, it can be represented by thefollowing formula.

T _(w) =Te·R _(TRQCVT) ·R _(AT) ·R _(FINAL)·η_(TOTAL)

Here, T_(w) denotes the wheel rim drive torque, Te the engine torque,R_(TRQCVT) the torque converter torque ratio, R_(AT) the gear ratio ofthe automatic transmission, R_(FINAL) the final gear ratio, andη_(TOTAL) the drive system efficiency.

The wheel rim drive torque likewise fluctuates during braking as well.In this case, because the braking force is proportional to the wheelcylinder pressure (in the case of normal braking in which control suchas ABS is not performed, essentially the master cylinder pressure), thebraking force on each wheel is estimated by multiplying a gain by themaster cylinder pressure.

When the wheel rim drive torque (or wheel rim braking torque) isestimated in the manner outlined above, filtering by the 0.5 Hz low-passfilter, which represents a region to the low-frequency end from afrequency region including stroke speed and sprung mass speed isperformed on the wheel rim drive torque by a low-pass filter 460, and astationary component is extracted. A pseudo-differential unit 461 thencalculates, through differentiation, a rate of change of the wheel rimdrive torque. In a rate of change decision unit 462, when the calculatedrate of change of wheel rim drive torque is equal to or greater than apre-established prescribed value dfthi, a deteriorated-estimationaccuracy signal is output to a wheel rim drive torque hold unit 463.Moreover, in order to prevent hunting during this decision, output ofthe deteriorated-estimation accuracy signal is halted in cases in whichthe differential is equal to or less than value obtained by multiplyingthe prescribed value dfthi by 0.8. In the wheel rim drive torque speedhold unit 463, for the time interval that the deteriorated-estimationaccuracy signal is being received, and for a prescribed intervalfollowing completion of reception (in the case of the first embodiment,a one-second interval), the deteriorated-accuracy signal is outputcontinuously. In so doing, a steady deteriorated-estimation accuracysignal can be output, even in cases in which the deteriorated-estimationaccuracy signal of the rate of change decision unit 462 repeatedly goesON/OFF. When, in the course of performing the above detection processes,deterioration in accuracy is detected in any one of these, thedeteriorated-estimation accuracy flag goes “ON,” and a control processappropriate at times of deterioration in accuracy is executed. Thecontrol process at times of deterioration in accuracy is describedbelow.

(Control Process at Times of Deterioration in Accuracy)

In cases in which the deteriorated-estimation accuracy flag is “ON,”specifically, when deterioration in accuracy of estimation of the strokespeed is detected, the deterioration estimation accuracy detectedcontrol unit 5 a outputs a “zero” as the engine orientation controlamount to the engine control unit 102. Additionally, in cases in whichthe deteriorated-estimation accuracy flag is “ON,” the deteriorationestimation accuracy detected control unit 5 a outputs a “zero” as thebrake orientation control amount to the brake control unit 202. At thistime, in the brake control unit 202, the brake orientation controlamount is gradually lowered in such a way that the brake orientationcontrol amount is lowered smoothly to zero over a fixed transitioninterval (for example, a one-second interval).

When brake pitch control for the purpose of holding the sprung masspitch speed to a low level is suddenly halted, the pitch speed, which upto this point was held to a low level, suddenly increases, which canresult in discomfort to the driver, as well as posing the risk ofincreased pitch behavior, or loss of ground contact load of the tires,leading to disturbances in vehicle behavior. By gradually lowering thebraking torque control amount in the manner discussed above, suddenincreases in the pitch speed can be minimized, and therefore discomfortto the driver can be reduced, and disturbances in vehicle behaviorminimized.

In the manner indicated above, at times that deterioration of estimationaccuracy of the stroke speed is detected, sprung mass vibration dampingcontrol by the engine 1 and the brakes 20 is suspended. In the firstembodiment, the stroke speed is estimated from fluctuations of wheelspeed in a prescribed frequency range, and sprung mass behavior controlusing the engine 1 and the brakes 20 is carried out according to thestroke speed; therefore, under conditions of deterioration in accuracyof estimation of the stroke speed, state estimation becomes difficult,and disturbances in sprung mass behavior due to degraded controllabilityare a concern. Therefore, in such cases, by suspending sprung massbehavior control by the engine 1 and the brakes 20, disturbances invehicle body orientation associated with deterioration of estimationaccuracy can be minimized, and stable vehicle body orientation can bemaintained.

In cases in which the deteriorated-estimation accuracy flag is “ON,” thedeterioration estimation accuracy detected control unit 5 a outputsdeteriorated-estimation accuracy-specific control signal (commandcurrent value) to the damping force control unit 35. FIG. 25 is acontrol block diagram showing a configuration of the deteriorationestimation accuracy detected control unit 5 a in the first embodiment. Avehicle speed calculating unit 501 inputs the deteriorated-estimationaccuracy flag, the vehicle speed VSP detected by the vehicle speedsensor 8, and the value of vehicle speed VSP observed in the previousone sampling cycle (one clock). In a case in which thedeteriorated-estimation accuracy flag is “OFF,” the vehicle speedcalculating unit 501 outputs to a delay element 502 the vehicle speedVSP detected by the vehicle speed sensor 8; and in a case in which thedeteriorated-estimation accuracy flag has gone “ON,” for a time intervaluntil the flag goes “OFF,” outputs to a damping coefficient setting unit503 the vehicle speed observed in the previous one sampling cycle,specifically, the vehicle speed observed just prior to whendeterioration of estimation accuracy of the stroke speed was detected.The delay element 502 delays the signal by one clock cycle. The dampingcoefficient setting unit 503 inputs the vehicle speed observed justprior to when deterioration of estimation accuracy of the stroke speedwas detected, the outside air temperature detected by the temperaturesensor 14, and the current driving mode, and outputs a deteriorationestimation accuracy damping coefficient k5. The method for establishingthe damping coefficient k will be discussed below. In a damping forcecontrol amount calculating unit 504, a control signal for the S/A 3 iscalculated based on the damping coefficient k5 and a predeterminedhypothetical stroke speed. Here, the hypothetical stroke speed is afixed value representing a stroke speed such that the damping force ofthe S/A 3 is comparable to the damping force of conventional shockabsorbers, for example, 0.1 m/s.

FIG. 26 is a descriptive diagram showing a method of setting adeterioration estimation accuracy damping coefficient, in the dampingcoefficient setting unit of the first embodiment. The dampingcoefficient k5 is basically a value proportional to the vehicle speed(the vehicle speed observed just prior to when deterioration ofestimation accuracy of the stroke speed was detected), and has a profilesuch that, for a given vehicle speed, the damping coefficient for thefront wheel side Fr is higher than the damping coefficient for the rearwheel side Rr. The value damping coefficient k5 also accords with thecontrol mode. In specific terms, the value is highest in the sport modeand the highway mode, and lowest in the comfort mode. In the standardmode, the value is an intermediate between the sport mode (the highwaymode) and the comfort mode. The damping coefficient k5 in the sport modeand the highway mode is a damping coefficient at the upper limit of thelevel at which vibration in the bounce region (3-6 Hz) is nottransmitted to passengers. In the case of the comfort mode, the dampingcoefficient setting unit 504 sets the damping coefficient in cases inwhich outside air temperature is outside a prescribed range (forexample, outside air temperature 5° C. or outside air temperature 30° C.or above) to a value higher than the damping coefficient in cases inwhich the outside air temperature is within the prescribed range (forexample, 5° C.<outside air temperature<30° C.) (the same dampingcoefficient as in the comfort mode).

In the manner indicated above, during times that deterioration ofestimation accuracy of the stroke speed is detected, vehiclespeed-sensitive control in which the damping force of the S/A 3 is setto a fixed damping force corresponding to the control mode is employed.The reason is that, at times of deterioration of estimation accuracy,despite the fact that a failure of a sensor signal or actuator drivinghas not occurred, state estimation is difficult nevertheless, andtherefore control in which stability is assigned greater weight,relative to ordinary times when estimation accuracy has notdeteriorated, is necessary. In vehicle speed-sensitive control, a fixeddamping coefficient is determined from the vehicle speed observed justprior to when deterioration of estimation accuracy was detected, and assuch is not dependent upon stroke speed having a high likelihood ofmistaken estimation, whereby, through a transition to a stable statefrom an unstable state dependent upon stroke speed, lowered steeringstability/ride comfort performance and destabilization of behavior canbe minimized.

In vehicle speed-sensitive control, the damping coefficient k5 isestablished based on the vehicle speed observed just prior to whendeterioration of estimation accuracy was detected, and a fixed dampingforce is determined based on the damping coefficient k5 and a prescribedhypothetical stroke speed (0.1 m/s). At this time, because the dampingcoefficient k5 is set to a value that is higher in association withhigher vehicle speed observed just prior to when deterioration ofestimation accuracy was detected, optimal damping force matched to thevehicle speed is obtained. That is, it is possible both to ensure ridecomfort in a low speed zone, and to ensure stable maneuvering in a highspeed zone.

Moreover, because the damping coefficient k5 assumes successively highervalues in the sport mode/highway mode, the standard mode, and thecomfort mode, in that order, a fixed damping force that is matched tothe control mode can be established. That is, in the sport mode andhighway mode, the damping force is high, giving priority to stablemaneuvering; in the comfort mode, the damping force is low, givingpriority to ride comfort; and in the standard mode, the damping force isintermediate, so that both stable maneuvering and ride comfort areachieved. Moreover, in each of the control modes, the fixed dampingforce of the front wheels exceeds the fixed damping force of the rearwheels, whereby nose diving can be minimized, an understeering tendencycan be adopted as the steering tendency, and stable turning behavior canbe ensured.

As shown above, during driving in the comfort mode, when deteriorationof estimation accuracy of the stroke speed is detected, and moreover theoutside air temperature is outside a prescribed range (for example,outside air temperature ≦5° C. or outside air temperature ≧30° C. orabove), the fixed damping force is set to a higher level than that incases in which the outside air temperature is outside the prescribedrange (for example, outside air temperature≦5° C. or outside airtemperature 30° C. or above). In other words, in cases in which loweredestimation precision has been detected during driving in a state of alow coefficient of friction between the tires and the road surface inthe comfort mode, the fixed damping force will be set to a higher levelthan that in cases of during driving in a state of a high coefficient offriction between the tires and the road surface. The reason is that, incases in which the outside air temperature is outside of the prescribedrange, it can be determined that the μ value of the road will be high,and when the outside air temperature is outside of the prescribed range,the coefficient of friction between the tires and the road surface canbe determined to be low, due to lowered gripping force of the tires.

In cases in which nose diving occurs due to deterioration of estimationaccuracy, producing change in a direction that reduces the groundcontact load of the tires of the rear wheels, when driving on a roadwith a high μ value such that the threshold limit of the gripping forceof the tires is high (the friction circle is large), there is a lowprobability of exceeding the threshold limit of the gripping force ofthe tires during turning; however, when driving on a road with a low μvalue such that the threshold limit of the gripping force of the tiresis low (the friction circle is small), there is a high probability ofexceeding the threshold limit of the gripping force of the tires duringturning. Particularly in the comfort mode, in which ride comfort haspriority, the fixed damping force assumes its lowest value as comparedwith the other three modes, and therefore the aforedescribed problembecomes quite noticeable. Therefore, in cases of deterioration ofestimation accuracy of stroke speed when driving on a road with a low μvalue in the comfort mode, there is a risk of lowered stability ofturning behavior, such as induced reverse steer or the like. Here,reverse steer refers to a change in the steering tendency, from anundersteer tendency to an oversteer tendency, during turning. For thisreason, according to the first embodiment, in the comfort mode, thefixed damping force when driving on a road with a low μ value is set toa higher level than the damping force when driving on a road with a highμ value. In so doing, the reduction in the ground contact load of therear tires can be minimized, and the occurrence of nose divingsuppressed, thereby minimizing the proclivity for the steering tendencyto become an oversteer tendency, and ensuring stable turning behavior.

On the other hand, in the sport mode, the highway mode, and the standardmode, the fixed damping force when driving on a road with a low μ valueis set to the same value as the damping force when driving on a roadwith a high μ value. The reason is that because the fixed damping forcein these three modes is greater than the fixed damping force in thecomfort mode, the probability of exceeding the threshold limit of thegripping force of the tires during turning when driving on a road with alow μ value is low. Moreover, in the sport mode and the highway mode,the fixed damping force equals a maximum fixed damping force at whichvibration producing bouncing motion is not transmitted to thepassengers, and therefore degradation of ride comfort can be minimized,and stability of maneuvering improved.

In the control signal conversion unit 35 c of the damping force controlunit 35, in a case in which a command current value (adeteriorated-estimation accuracy-specific command current value) hasbeen output by the deterioration estimation accuracy detected controlunit 5 a, the deteriorated-estimation accuracy-specific command currentvalue is output to the S/A 3, in place of a command current value(normal command current value) based on the stroke speed and the dampingcoefficient reconciled by the damping coefficient reconciling unit 35 b.At this time, the command current value is changed gradually, so as tosmoothly transition the command current value from the current commandcurrent value, to the deteriorated-estimation accuracy-specific commandcurrent value, doing so over a prescribed transition time. Here, thistransition time is a time equivalent at least to the cycle of the sprungmass resonance (1.2 Hz) or less (for example, 0.5 Hz), for example, aone-second interval. During switching of the command current value forthe S/A 3 from a normal command current value based on the stroke speed,to the deteriorated-estimation accuracy-specific command current valuebased on vehicle speed, in cases in which the command current value forthe S/A 3 exceeds the difference of the two command current values,there is a risk of sudden change of the damping force of the S/A,disturbing the vehicle body orientation. By gradually changing thecommand current value in the aforedescribed manner to limit the extentof fluctuation of the damping force as indicated above, disturbance ofthe vehicle body orientation at times of deterioration of estimationaccuracy can be minimized.

As described above, the first embodiment affords the following effects.

(1) The system is provided with S/A 3 (actuators) for performing sprungmass vibration damping control, wheel speed sensors 5 (wheel speeddetection means) for detecting wheel speed, a third driving stateestimating means 32 (sprung mass state estimating means) for estimatingthe sprung mass state based on information about wheel speeds detectedby the wheel speed sensors 5 in a prescribed frequency region, an S/Acontroller 3 a (actuator orientation control means) for controlling theS/A 3 (actuators) so as to bring the estimated sprung mass state to atarget sprung mass state, and an estimation accuracy deteriorationdetection unit 4 a (estimation accuracy deterioration detection means)for detecting deterioration of estimation accuracy by the third drivingstate estimating means 32, the control executed by the actuatororientation control means in cases in which deterioration of estimationaccuracy has been detected by the estimation accuracy deteriorationdetection unit 4 a being more limited, as compared with the case inwhich the estimation accuracy has not been lowered. Therefore,deterioration of estimation accuracy of the sprung mass state can bedetected, and situations in which control continues while estimationaccuracy remains at a lowered level can be avoided. Additionally,because at times of deterioration of estimation accuracy, thedeterioration estimation accuracy detected control unit 5 a performsdamping force control that is more limited than that prior to loweringof estimation accuracy, instances of erroneous control can be minimized,and stable vehicle orientation achieved.

(2) The system is provided with a deterioration estimation accuracydetected control unit 5 a (braking control means) for transitioning thedamping force of the S/A 3 to a fixed damping force in accordance withvehicle speed (vehicle state quantity) observed prior to whendeterioration of estimation accuracy was detected by the estimationaccuracy deterioration detection unit 4 a. Therefore, deterioration ofestimation accuracy of the sprung mass state can be detected, andsituations in which control continues while estimation accuracy remainsat a lowered level can be avoided. Additionally, at times ofdeterioration of estimation accuracy, the deterioration estimationaccuracy detected control unit 5 a transitions to a fixed damping forcein accordance with the vehicle speed as a vehicle state quantityobserved prior to detection of deterioration of estimation accuracy,whereby instances of erroneous control can be minimized, and stablevehicle orientation achieved.

(3) The S/A controller 3 a has a highway mode, a sport mode, a standardmode, and a comfort mode (plurality of control modes) for whichdifferent damping control ranges have been established for a givenstroke speed, and the deterioration estimation accuracy detected controlunit 5 a transitions to a fixed damping force in accordance with theprevailing control mode at the time that deterioration of estimationaccuracy was detected by the estimation accuracy deterioration detectionunit 4 a. For example, in a case in which deterioration of estimationaccuracy was detected during driving in the comfort mode, there can beenvisioned a scenario in which fixing the damping coefficient at a lowvalue would make it difficult to ensure sufficient stability on the partof a vehicle state. In such cases, stability can be ensured by fixingthe damping force at a higher level than the damping force establishedin the comfort mode.

(4) When the deterioration estimation accuracy detected control unit 5 atransitions the damping force of the S/A 3 to a fixed damping force, thechange is gradual, taking place over a transition time equivalent to thefrequency of the sprung mass resonance frequency or less (in the firstembodiment, one second). Therefore, destabilization of the vehicle inassociation with fluctuations during transition to the fixed dampingforce can be avoided, and vehicle stability can be ensured.

(5) The system is provided with a vehicle speed sensor 8, a steeringangle sensor 7, an integrated sensor 6 for detecting an actual vehiclestate of actual yaw rate and/or actual lateral acceleration, and avehicle motion state estimating unit 430 (vehicle motion stateestimating means) for inputting the detected wheel speed and steeringangle, and estimating a vehicle state of yaw rate and/or lateralacceleration based on a plan view model of the vehicle, wherein theestimation accuracy deterioration detection unit 4 a (estimationaccuracy deterioration detection means) compares the actual vehiclestate detected by the vehicle motion state estimating unit 430, and theestimated vehicle state estimated by the vehicle motion state estimatingmeans, in terms of information to the low-frequency end from aprescribed frequency range in which sprung behavior is observed, and incases of divergence by a prescribed level or above between the two setsof information, detects that the estimation accuracy of the firstdriving state estimating unit 100, the second driving state estimatingunit 200, and the third driving state estimating unit 32 hasdeteriorated. Therefore, deterioration of estimation accuracy of thesprung mass state can be detected, and situations in which controlcontinues while estimation accuracy remains at a lowered level can beavoided. Additionally, through limiting of functionality by thedeterioration estimation accuracy detected control unit 5 a at times ofdeterioration of estimation accuracy, instances of erroneous control canbe minimized, and stable vehicle orientation achieved.

(6) The deterioration estimation accuracy detected control unit 5 atransitions to a fixed damping force in accordance with the vehiclespeed observed just prior to detection of deterioration of estimationaccuracy by the estimation accuracy deterioration detection unit 4 a.Therefore, a fixed damping force can be established based on the vehiclespeed actually used in each control mode, rather than the vehicle speedat deterioration of estimation accuracy, and vehicle stability can beimproved.

(7) As actuators, the system has an engine 1 (vehicle power source) andbrakes 20 (friction brakes), the actuator control means has an enginecontroller 1 a and a brake controller 2 a, and in cases in whichdeterioration of estimation accuracy has been detected by the estimationaccuracy deterioration detection unit 4 a, the deterioration estimationaccuracy detected control unit 5 a suspends control by the enginecontroller 1 a and the brake controller 2 a. Specifically, when anactuator that contributes to forward/reverse acceleration of thevehicle, such as braking/drive torque, performs torque control in aforward/reverse direction using erroneous information or information oflow accuracy, there is a risk of creating unintentional accelerationbearing no relationship to the sprung mass state. In contrast to this,when control affecting braking/drive torque in a forward/reversedirection is suspended, the risk of discomfort to the driver can beavoided.

(8) The system has a vehicle speed sensor 8 (vehicle speed detectingmeans) for detecting the vehicle speed; and the deterioration estimationaccuracy detected control unit 5 a transitions to a higher fixed dampingforce, in association with higher vehicle speed as a vehicle statequantity. Therefore, stable damping force with respect to vehicle speedcan be ensured. Moreover, because higher fixed damping force isassociated with higher vehicle speed, better vehicle stability can beensured.

(9) Whereas the first embodiment showed an example in which vehiclespeed is employed as a vehicle state quantity, because the system has anintegrated sensor 7 for detecting the vehicle yaw rate (yaw ratedetecting means), the deterioration estimation accuracy detected controlunit 5 a could instead transition to a higher fixed damping force, inassociation with higher yaw rate as a vehicle state quantity. Therefore,by fixing the damping force at a high level in cases in which thevehicle is in a turning motion state, vehicle stability can be ensured.There is no limitation to yaw rate, and it would be acceptable to employlateral acceleration or the like, or to estimate various other vehiclestate quantities from steering angle and vehicle speed information, anduse these values to establish fixed damping forces. In the case ofemploying lateral acceleration, higher fixed damping force would beestablished in association with greater lateral acceleration.

(10) The system has wheel speed sensors 5 for detecting wheel speed, andis provided with an S/A controller 3 a (control device) which estimatesthe sprung mass state based on information in a prescribed frequencyrange detected by the wheel speed sensors 5, and controls the S/A 3 soas to bring this sprung mass state to a target sprung mass state, andwhich, in the event of deterioration of estimation accuracy of thesprung mass state, performs more limited damping force control, ascompared with that when estimation accuracy has not deteriorated.Therefore, deterioration of estimation accuracy of the sprung mass statecan be detected, and situations in which control continues whileestimation accuracy remains at a lowered level can be avoided.Additionally, through limiting of damping force control by thedeterioration estimation accuracy detected control unit 5 a at times ofdeterioration of estimation accuracy, instances of erroneous control canbe minimized, and stable vehicle orientation achieved.

(11) The system has wheel speed sensors 5 for detecting wheel speed, andan S/A controller 3 a (control device) estimates the sprung mass statebased on information in a prescribed frequency range detected by thewheel speed sensors 5, and controls the S/A 3 so as to bring this sprungmass state to a target sprung mass state, as well as, in the event ofdeterioration of estimation accuracy of the sprung mass state,performing damping force control of the damping force of the S/A 3 to amore limited extent, as compared with that when estimation accuracy hasnot deteriorated. Therefore, deterioration of estimation accuracy of thesprung mass state can be detected, and situations in which controlcontinues while estimation accuracy remains at a lowered level can beavoided. Additionally, through limiting the extent of damping forcecontrol by the deterioration estimation accuracy detected control unit 5a at times of deterioration of estimation accuracy, instances oferroneous control can be minimized, and stable vehicle orientationachieved.

1. A vehicle control device comprising: a variable-damping force shockabsorber capable of varying a damping force at which sprung massvibration damping control is carried out; a wheel speed sensor thatdetects wheel speed; a sprung mass state estimating unit that estimatessprung mass state based on information in a prescribed frequency rangeof wheel speed detected by the wheel speed sensor; an actuatororientation control unit that controls the variable-damping force shockabsorber so as to bring the estimated sprung mass state to a targetsprung mass state; an estimation accuracy deterioration detection unitthat detects deterioration in estimation accuracy performed by thesprung mass state estimating unit and that estimates an estimated wheelrim braking/drive torque to determine that the estimation accuracy hasdeteriorated when a rate of change of a stationary component extractedfrom components of wheel rim braking/drive torque acting on a wheel isdetected to equal or exceed a prescribed value, with the stationarycomponent of the estimated wheel rim braking/drive torque lying in alow-frequency end of a frequency region that includes a stroke speed anda sprung mass speed; and a limited control unit that implements controlperformed by the actuator orientation control unit to a more limitedextent upon detecting deterioration of the estimation accuracy by theestimation accuracy deterioration detection means unit as compared towhen estimation accuracy has not deteriorated.
 2. The vehicle controldevice according to claim 1, wherein the limited control unittransitions the damping force of the variable-damping force shockabsorber to a fixed damping force corresponding to a vehicle statequantity observed prior to detection of deterioration of estimationaccuracy by the estimation accuracy detecting unit.
 3. The vehiclecontrol device according to claim 1, wherein the actuator orientationcontrol unit has a plurality of control modes in which different dampingforce control ranges are established for a given stroke speed; and thelimited control unit transitions to a fixed damping force correspondingto the control mode prevailing at a time that deterioration ofestimation accuracy was detected by the estimation accuracy detectingunit.
 4. The vehicle control device according to claim 1, wherein thelimited control unit, when transitioning the damping force of thevariable-damping force shock absorber to a fixed damping force, bringsabout gradual change over a transition period during a frequency becomesequal to or less than a sprung mass resonance frequency.
 5. The vehiclecontrol device according to claim 1, further comprising a wheel speedsensor that detects wheel speed; a steering angle sensor that detectssteering angle; a vehicle state sensor that detects an actual vehiclestate constituted by an actual yaw rate and/or actual lateralacceleration of the vehicle; and a vehicle state estimating unit thatinputs the detected wheel speed and steering angle, and that estimates avehicle state constituted by a yaw rate and/or lateral acceleration,based on a plan view model of the vehicle; the estimation accuracydeterioration detection unit being constituted by comparing the actualvehicle state detected by the vehicle state detecting unit, and theestimated vehicle state estimated by the vehicle state estimating unit,in means of information to the low-frequency end from the prescribedfrequency range in which sprung behavior is observed, and in cases ofdivergence by a prescribed level or above between two sets ofinformation, detecting that the estimation accuracy of the sprung massstate has deteriorated.
 6. The vehicle control device according to claim1, wherein the limited control unit transitions to a fixed damping forcecorresponding to vehicle speed immediately prior to when deteriorationof estimation accuracy was detected by the estimation accuracy detectingunit.
 7. The vehicle control device according to claim 1, furthercomprising a vehicle power source and a friction brake for performingsprung mass vibration damping control; a power source orientationcontrol unit that controls torque of the power source so as to bring theestimated sprung mass state to a target sprung mass state; and a brakeorientation control unit that controls torque of the friction brake soas to bring the estimated sprung mass state to a target sprung massstate; the limited control unit, when deterioration of estimationaccuracy has been detected by the estimation accuracy detecting unit,suspending control performed by the power source orientation controlunit and the brake orientation control unit.
 8. The vehicle controldevice according to claim 1, further comprising a wheel speed sensorthat detects wheel speed; the limited control unit transitioning to ahigher fixed damping force in association with higher vehicle speed as avehicle state quantity.
 9. The vehicle control device according to claim1, further comprising a yaw rate sensor that detects a yaw rate of thevehicle, the limited control unit transitioning to a higher fixeddamping force in association with higher yaw rate as the vehicle statequantity.
 10. A control device for a vehicle, comprising: a sensor fordetecting wheel speed, and a controller for estimating sprung mass statebased on information in a prescribed frequency range of wheel speeddetected by the sensor, for controlling a variable-damping force shockabsorber so as to bring the estimated sprung mass state to a targetsprung mass state, for estimating wheel rim braking/drive torque actingon a wheel and for determining that the estimation accuracy of thesprung mass state has deteriorated when a rate of change of a stationarycomponent extracted from components of wheel rim braking/drive torqueacting on a wheel is detected to equal or exceed a prescribed value,with the stationary component of the estimated wheel rim braking/drivetorque lying in a low-frequency end of a frequency region that includesa stroke speed and a sprung mass speed, and, for controlling thevariable-damping force shock absorber to a more limited extent when theestimation accuracy of the sprung mass state has deteriorated than whenthe estimation accuracy has not deteriorated.
 11. A control method forcontrolling a vehicle, the method comprising: detecting wheel speedusing a sensor; and estimating, using a controller a sprung mass statebased on information in a prescribed frequency range of wheel speeddetected by the sensor; controlling, using the controller, avariable-damping force shock absorber so as to bring the estimatedsprung mass state to a target sprung mass state; estimating, using thecontroller, wheel rim braking/drive torque acting on a wheel todetermine that the estimation accuracy has deteriorated when a rate ofchange of a stationary component extracted from components of wheel rimbraking/drive torque acting on a wheel is detected to equal or exceed aprescribed value, with the stationary component of the estimated wheelrim braking/drive torque lying in a low-frequency end of a frequencyregion that includes a stroke speed and a sprung mass speed, and whenthe estimation accuracy of the sprung mass state has deteriorated,controlling, using the controller, the variable-damping force shockabsorber to a more limited extent than when the estimation accuracy hasnot deteriorated.